U.S. patent application number 12/469318 was filed with the patent office on 2009-11-26 for variable displacement vane pump.
This patent application is currently assigned to HITACHI, LTD.. Invention is credited to Hideaki Ohnishi, Koji Saga, Yasushi Watanabe.
Application Number | 20090291000 12/469318 |
Document ID | / |
Family ID | 41342261 |
Filed Date | 2009-11-26 |
United States Patent
Application |
20090291000 |
Kind Code |
A1 |
Saga; Koji ; et al. |
November 26, 2009 |
VARIABLE DISPLACEMENT VANE PUMP
Abstract
In a variable displacement vane pump employing a cam ring, a
pump rotor, and a plurality of vanes, each slidably fitted into the
rotor, a biasing member is provided to force the cam ring in a
direction that a geometric center of an inner peripheral surface of
the cam ring and a rotation center of the rotor are spaced apart
from each other. A force, by which the cam ring can be oscillated
against the biasing member in accordance with a buildup of a
pressure in a pump discharge portion, acts on the inner peripheral
surface of the cam ring.
Inventors: |
Saga; Koji; (Ebina-shi,
JP) ; Watanabe; Yasushi; (Aiko-gun, JP) ;
Ohnishi; Hideaki; (Atsugi-shi, JP) |
Correspondence
Address: |
FOLEY AND LARDNER LLP;SUITE 500
3000 K STREET NW
WASHINGTON
DC
20007
US
|
Assignee: |
HITACHI, LTD.
|
Family ID: |
41342261 |
Appl. No.: |
12/469318 |
Filed: |
May 20, 2009 |
Current U.S.
Class: |
417/364 ;
418/24 |
Current CPC
Class: |
F04C 14/226 20130101;
F04C 2/3564 20130101; F04C 2/3442 20130101 |
Class at
Publication: |
417/364 ;
418/24 |
International
Class: |
F04B 17/05 20060101
F04B017/05; F04C 2/344 20060101 F04C002/344 |
Foreign Application Data
Date |
Code |
Application Number |
May 22, 2008 |
JP |
2008-133889 |
Claims
1. A variable displacement vane pump comprising: a rotor driven by
an internal combustion engine; a cam ring configured to accommodate
therein the rotor and further configured to oscillate about a
fulcrum of oscillating motion along two axially opposed sidewalls
facing both sides of the cam ring respectively; a plurality of
vanes, each of which is fitted into the rotor to slide from the
rotor toward an inner peripheral surface of the cam ring and set to
be kept in abutted-engagement with the inner peripheral surface of
the cam ring, the vanes being configured to define a plurality of
working chambers in cooperation with an outer peripheral surface of
the rotor, the inner peripheral surface of the cam ring, and the
two axially opposed sidewalls; a biasing member configured to force
the cam ring in a direction that a geometric center of the inner
peripheral surface of the cam ring and a rotation center of the
rotor are spaced apart from each other; and an inlet portion and a
discharge portion both formed in at least one of the two axially
opposed sidewalls, the inlet portion being configured to open into
a first group of working chambers of the plurality of working
chambers so as to extend over the first group of working chambers
within an area where volumes of the first group of working chambers
increase, and the discharge portion being configured to open into a
second group of working chambers of the plurality of working
chambers so as to extend over the second group of working chambers
within an area where volumes of the second group of working
chambers decrease, wherein a force, by which the cam ring can be
oscillated against the biasing member in accordance with a buildup
of a pressure in the discharge portion, acts on the inner
peripheral surface of the cam ring.
2. The variable displacement vane pump as claimed in claim 1,
wherein: a pressure, acting on an outer peripheral surface of the
cam ring, is lower than the pressure in the discharge portion.
3. The variable displacement vane pump as claimed in claim 2,
wherein: atmospheric pressure is applied on the outer peripheral
surface of the cam ring.
4. The variable displacement vane pump as claimed in claim 2,
wherein: the discharge portion comprises grooves formed in the two
axially opposed sidewalls; the cam ring has a communication hole,
which is formed in the cam ring so as to axially penetrate the cam
ring and through which the discharge portions formed in the
respective sidewalls are communicated with each other; and working
fluid is discharged through a grooved portion of at least one of
the discharge portions, the grooved portion being configured to be
substantially conformable to a shape of the communication hole of
the cam ring, via a discharge hole to an exterior space.
5. The variable displacement vane pump as claimed in claim 4,
wherein: a fluid-flow passage cross-sectional area of the
communication hole is dimensioned to be greater than or equal to a
fluid-flow passage cross-sectional area of the discharge hole.
6. The variable displacement vane pump as claimed in claim 4,
wherein: the communication hole is formed into a circular-arc shape
whose center is the fulcrum of oscillating motion of the cam
ring.
7. The variable displacement vane pump as claimed in claim 4,
wherein: the communication hole is configured to displace without
any change in an opening area of the communication hole opening
into the discharge hole, during oscillating motion of cam ring
5.
8. The variable displacement vane pump as claimed in claim 1,
wherein: a radial wall thickness of a part of the cam ring,
overlapping with the inlet portion and the discharge portion, is
dimensioned to be greater than a radial wall thickness of the other
part of the cam ring.
9. The variable displacement vane pump as claimed in claim 1,
wherein: the biasing member comprises a first biasing member that
permanently forces the cam ring, and a second biasing member that
exerts a biasing force on the cam ring only when the cam ring
oscillates a predetermined distance, which distance is greater than
or equal to a predetermined angular displacement.
10. The variable displacement vane pump as claimed in claim 1,
wherein: the cam ring has a through hole into which a pin, serving
as the fulcrum of oscillating motion of the cam ring, is
inserted.
11. The variable displacement vane pump as claimed in claim 1,
wherein: the sidewalls are made of an aluminum alloy material,
whereas the cam ring is made of an iron-based material.
12. The variable displacement vane pump as claimed in claim 1,
wherein: working fluid, which is discharged from the discharge
portion, is lubricating oil, which lubricating oil is supplied to
moving engine parts of the internal combustion engine, the working
fluid is also used as a power source for a variable valve actuation
system configured to vary a valve characteristic of the internal
combustion engine.
13. A variable displacement vane pump comprising: a rotor driven in
synchronism with rotation of an internal combustion engine; a cam
ring configured to accommodate the rotor in an inner peripheral
surface of the cam ring and further configured to oscillate about a
fulcrum of oscillating motion between two axially opposed sidewalls
facing both sides of the cam ring respectively; a plurality of
vanes, each of which is fitted into the rotor to slide from an
outer peripheral surface of the rotor toward the inner peripheral
surface of the cam ring, the vanes being configured to define a
plurality of working chambers in cooperation with the outer
peripheral surface of the rotor, the inner peripheral surface of
the cam ring, and the two axially opposed sidewalls; a biasing
member configured to force the cam ring in a direction that a
volume difference between a volume of the largest working chamber
of the plurality of working chambers and a volume of the smallest
working chamber of the plurality of working chambers increases; and
an inlet portion and a discharge portion both formed in at least
one of the two axially opposed sidewalls, the inlet portion being
configured to open into a first group of working chambers of the
plurality of working chambers so as to extend over the first group
of working chambers within an area where volumes of the first group
of working chambers increase, and the discharge portion being
configured to open into a second group of working chambers of the
plurality of working chambers so as to extend over the second group
of working chambers within an area where volumes of the second
group of working chambers decrease, wherein the fulcrum of
oscillating motion of the cam ring is laid out to be offset in a
biasing direction of the biasing member within an opening range of
the discharge portion.
14. The variable displacement vane pump as claimed in claim 13,
wherein: an approximately uniform pressure is applied around an
entire circumference of an outer peripheral surface of the cam
ring.
15. The variable displacement vane pump as claimed in claim 13,
wherein: the biasing member is located outside of an outer
periphery of the cam ring and laid out to be offset toward the
fulcrum of oscillating motion with respect to a geometric center of
the inner peripheral surface of the cam ring.
16. The variable displacement vane pump as claimed in claim 13,
wherein: the biasing member comprises a plurality of springs; the
cam ring is forced only by one of the plurality of springs when an
oscillated amount of the cam ring is less than or equal to a
predetermined threshold value; and the cam ring is forced by the
plurality of springs when the oscillated amount of the cam ring
exceeds the predetermined threshold value.
17. The variable displacement vane pump as claimed in claim 16,
wherein: working fluid, which is discharged from the discharge
portion, is supplied to a variable valve timing control system of
the internal combustion engine; the variable valve timing control
system is configured to hold an engine valve timing at a locked
state during a startup period of the engine, and further configured
to release the locked state of the valve timing by a pressure of
the working fluid discharged from the discharge portion after the
engine has been started up, so as to permit the valve timing to be
varied to a desired valve timing; and a pressure level of the
working-fluid pressure, at which the locked state of the valve
timing is released, is set to be lower than a pressure level of the
working-fluid pressure, at which the cam ring begins to operate
against a biasing force of the biasing member.
18. The variable displacement vane pump as claimed in claim 17,
wherein: the variable valve timing control system is configured to
operate by the pressure of the working fluid, discharged from the
discharge portion, and further configured to be able to operate in
a state where the cam ring is forced only by one of the plurality
of springs.
19. A variable displacement vane pump comprising: a rotor rotated
by a drive source; a cam ring configured to accommodate therein the
rotor and further configured to oscillate about a fulcrum of
oscillating motion, while being kept in sliding-contact with two
axially opposed sidewalls facing both sides of the cam ring
respectively; a plurality of vanes, each of which is fitted into
the rotor to slide from the rotor toward an inner peripheral
surface of the cam ring, the vanes being configured to define a
plurality of working chambers in cooperation with an outer
peripheral surface of the rotor, the inner peripheral surface of
the cam ring, and the two axially opposed sidewalls; a biasing
member configured to force the cam ring in a biasing direction that
a rate of change of a volume of each of the plurality of working
chambers increases; and an inlet portion and a discharge portion
both formed in at least one of the two axially opposed sidewalls,
the inlet portion being configured to open into a first group of
working chambers of the plurality of working chambers so as to
extend over the first group of working chambers within an area
where volumes of the first group of working chambers increase, and
the discharge portion being configured to open into a second group
of working chambers of the plurality of working chambers so as to
extend over the second group of working chambers within an area
where volumes of the second group of working chambers decrease,
wherein an integral .intg.S2dt of a second segmented
pressure-receiving area of the inner peripheral surface of the cam
ring, extending in the biasing direction of the biasing member with
respect to the fulcrum of oscillating motion, for a given cycle, is
less than an integral .intg.S1dt of a first segmented
pressure-receiving area of the inner peripheral surface of the cam
ring, extending in the direction opposite to the biasing direction
of the biasing member with respect to the fulcrum of oscillating
motion, for the given cycle.
20. The variable displacement vane pump as claimed in claim 19,
wherein: the biasing member is configured such that a biasing force
per unit oscillated amount increases, as an oscillated amount of
the cam ring increases.
Description
TECHNICAL FIELD
[0001] The present invention relates to a variable displacement
vane pump whose discharge can be varied by changing an eccentricity
of a geometric center of a cylinder bore of a cam ring with respect
to the axis of rotation of a vane rotor.
BACKGROUND ART
[0002] In recent years, there have been proposed and developed
various variable displacement vane pumps capable of varying a
discharge of working fluid, usually expressed as a fluid flow rate
per one revolution of a vane-pump rotor. One such variable
displacement vane pump has been disclosed in Japanese Patent
Provisional Publication No. 05-79469 (hereinafter is referred to as
"JP5-079469"). The variable displacement vane pump disclosed in
JP5-079469, has a control oil chamber defined between the inner
periphery of a vane-pump housing and the outer periphery of a cam
ring and partitioned by a cam-ring pivot pin fixedly connected to
the pump housing and a seal member attached to the outer periphery
of the cam ring. The eccentricity of the cam ring with respect to
the vane rotor, exactly, the distance from the axis of rotation of
the vane rotor to the geometric center of the cylinder bore of the
cam ring, can be controlled or adjusted by varying a hydraulic
pressure supplied into the control oil chamber, thereby varying the
discharge of the vane pump.
[0003] However, to produce an oscillating motion of the cam ring
pivoted to the pivot pin, the variable displacement vane pump of
JP5-079469 requires the previously-noted oil control chamber
defined between the pump-housing inner periphery and the cam-ring
outer periphery and partitioned by the pivot pin and the seal
member. The disadvantages of the variable displacement vane pump of
JP5-079469 are the difficulty of reducing the number of component
parts constructing the variable displacement vane pump assembly and
the increased vane pump manufacturing costs.
[0004] Thus, it would be desirable to provide a variable
displacement vane pump of reduced number of component parts.
SUMMARY OF THE INVENTION
[0005] It is, therefore, in view of the previously-described
disadvantages of the prior art, an object of the invention to
provide a variable displacement vane pump, which is configured to
realize reduced number of component parts without the need for a
control oil chamber, defined between a pump-housing inner periphery
and a cam-ring outer periphery and partitioned by a plurality of
components, namely, a cam-ring pivot pin and a seal member.
[0006] In order to accomplish the aforementioned and other objects
of the present invention, a variable displacement vane pump
comprises a rotor driven by an internal combustion engine, a cam
ring configured to accommodate therein the rotor and further
configured to oscillate about a fulcrum of oscillating motion along
two axially opposed sidewalls facing both sides of the cam ring
respectively, a plurality of vanes, each of which is fitted into
the rotor to slide from the rotor toward an inner peripheral
surface of the cam ring and set to be kept in abutted-engagement
with the inner peripheral surface of the cam ring, the vanes being
configured to define a plurality of working chambers in cooperation
with an outer peripheral surface of the rotor, the inner peripheral
surface of the cam ring, and the two axially opposed sidewalls, a
biasing member configured to force the cam ring in a direction that
a geometric center of the inner peripheral surface of the cam ring
and a rotation center of the rotor are spaced apart from each
other, and an inlet portion and a discharge portion both formed in
at least one of the two axially opposed sidewalls, the inlet
portion being configured to open into a first group of working
chambers of the plurality of working chambers so as to extend over
the first group of working chambers within an area where volumes of
the first group of working chambers increase, and the discharge
portion being configured to open into a second group of working
chambers of the plurality of working chambers so as to extend over
the second group of working chambers within an area where volumes
of the second group of working chambers decrease, wherein a force,
by which the cam ring can be oscillated against the biasing member
in accordance with a buildup of a pressure in the discharge
portion, acts on the inner peripheral surface of the cam ring.
[0007] According to another aspect of the invention, a variable
displacement vane pump comprises a rotor driven in synchronism with
rotation of an internal combustion engine, a cam ring configured to
accommodate the rotor in an inner peripheral surface of the cam
ring and further configured to oscillate about a fulcrum of
oscillating motion between two axially opposed sidewalls facing
both sides of the cam ring respectively, a plurality of vanes, each
of which is fitted into the rotor to slide from an outer peripheral
surface of the rotor toward the inner peripheral surface of the cam
ring, the vanes being configured to define a plurality of working
chambers in cooperation with the outer peripheral surface of the
rotor, the inner peripheral surface of the cam ring, and the two
axially opposed sidewalls, a biasing member configured to force the
cam ring in a direction that a volume difference between a volume
of the largest working chamber of the plurality of working chambers
and a volume of the smallest working chamber of the plurality of
working chambers increases, and an inlet portion and a discharge
portion both formed in at least one of the two axially opposed
sidewalls, the inlet portion being configured to open into a first
group of working chambers of the plurality of working chambers so
as to extend over the first group of working chambers within an
area where volumes of the first group of working chambers increase,
and the discharge portion being configured to open into a second
group of working chambers of the plurality of working chambers so
as to extend over the second group of working chambers within an
area where volumes of the second group of working chambers
decrease, wherein the fulcrum of oscillating motion of the cam ring
is laid out to be offset in a biasing direction of the biasing
member within an opening range of the discharge portion.
[0008] According to a further aspect of the invention, a variable
displacement vane pump comprises a rotor rotated by a drive source,
a cam ring configured to accommodate therein the rotor and further
configured to oscillate about a fulcrum of oscillating motion,
while being kept in sliding-contact with two axially opposed
sidewalls facing both sides of the cam ring respectively, a
plurality of vanes, each of which is fitted into the rotor to slide
from the rotor toward an inner peripheral surface of the cam ring,
the vanes being configured to define a plurality of working
chambers in cooperation with an outer peripheral surface of the
rotor, the inner peripheral surface of the cam ring, and the two
axially opposed sidewalls, a biasing member configured to force the
cam ring in a biasing direction that a rate of change of a volume
of each of the plurality of working chambers increases, and an
inlet portion and a discharge portion both formed in at least one
of the two axially opposed sidewalls, the inlet portion being
configured to open into a first group of working chambers of the
plurality of working chambers so as to extend over the first group
of working chambers within an area where volumes of the first group
of working chambers increase, and the discharge portion being
configured to open into a second group of working chambers of the
plurality of working chambers so as to extend over the second group
of working chambers within an area where volumes of the second
group of working chambers decrease, wherein an integral .intg.S2dt
of a second segmented pressure-receiving area of the inner
peripheral surface of the cam ring, extending in the biasing
direction of the biasing member with respect to the fulcrum of
oscillating motion, for a given cycle, is less than an integral
.intg.S1dt of a first segmented pressure-receiving area of the
inner peripheral surface of the cam ring, extending in the
direction opposite to the biasing direction of the biasing member
with respect to the fulcrum of oscillating motion, for the given
cycle.
[0009] The other objects and features of this invention will become
understood from the following description with reference to the
accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0010] FIG. 1 is a cross-sectional view illustrating a variable
valve timing control (VTC) system employing a
hydraulically-operated vane-type timing variator (simply, a
hydraulic actuator) to which a variable displacement vane pump of
the first embodiment is applied.
[0011] FIG. 2 is a front elevation view of the hydraulic actuator
of the VTC device to which the variable displacement vane pump of
the first embodiment is applied.
[0012] FIG. 3 is a disassembled view of the variable displacement
vane pump of the first embodiment.
[0013] FIG. 4 is a front elevation view of a pump housing of the
variable displacement vane pump of the first embodiment.
[0014] FIG. 5 is a front elevation view of the variable
displacement vane pump of the first embodiment in an initial
setting state.
[0015] FIG. 6 is a cross section of the vane pump, taken along the
line E-E of FIG. 5.
[0016] FIG. 7 is an explanatory view illustrating the initial
setting position (the maximum-eccentricity angular position) of the
cam ring of the variable displacement vane pump of the first
embodiment.
[0017] FIG. 8 is an explanatory view illustrating the
minimum-eccentricity angular position of the cam ring of the
variable displacement vane pump of the first embodiment.
[0018] FIG. 9 is a time chart (a characteristic diagram)
illustrating a change in each of a first pressure-receiving area S1
and a second pressure-receiving area S2 of the cam-ring inner
peripheral surface of the variable displacement vane pump of the
first embodiment.
[0019] FIG. 10 is a front elevation view of the variable
displacement vane pump of the first embodiment in an
intermediate-eccentricity holding state where the cam-ring
eccentricity is held at a substantially intermediate value between
the maximum and minimum eccentricities.
[0020] FIG. 11 is a front elevation view of the variable
displacement vane pump of the first embodiment in the
minimum-eccentricity state.
[0021] FIG. 12 is a characteristic diagram showing the relationship
between a displacement of a biasing member (i.e., a cam-ring
oscillated angle) and a spring load in the variable displacement
vane pump of the first embodiment.
[0022] FIG. 13 is an engine-speed versus pump-discharge-pressure
characteristic diagram.
[0023] FIGS. 14A-14C are explanatory views illustrating the
relationship among the angular position of a vane member, the
position of a retractable lock piston, and the axial position of an
axially-slidable valve spool, in the VTC system, to which the
variable displacement vane pump of the first embodiment is applied,
in an engine stopped state.
[0024] FIGS. 15A-15C are explanatory views illustrating the
relationship among the angular position of the vane member, the
position of the retractable lock piston, and the axial position of
the valve spool, in the VTC system, to which the variable
displacement vane pump of the first embodiment is applied, during
an engine startup period.
[0025] FIGS. 16A-16C are explanatory views illustrating the
relationship among the angular position of the vane member, the
position of the retractable lock piston, and the axial position of
the valve spool, in the VTC system, to which the variable
displacement vane pump of the first embodiment is applied, at
middle engine-speed operation.
[0026] FIG. 17 is an engine-speed versus pump-discharge-pressure
characteristic diagram in a variable displacement vane pump of a
comparative example.
[0027] FIG. 18 is a disassembled view of a variable displacement
vane pump of the second embodiment.
[0028] FIG. 19 is a front elevation view of a pump housing of the
variable displacement vane pump of the second embodiment.
[0029] FIG. 20 is a front elevation view of the variable
displacement vane pump of the second embodiment in the initial
setting state.
[0030] FIG. 21 is a cross section taken along the line F-F of FIG.
20.
[0031] FIG. 22 is a front elevation view of the variable
displacement vane pump of the second embodiment in a holding state
where the cam-ring eccentricity is held at a substantially
intermediate value between the maximum and minimum
eccentricities.
[0032] FIG. 23 is a front elevation view of the variable
displacement vane pump of the second embodiment in the
minimum-eccentricity state.
[0033] FIG. 24 is a front elevation view of a pump housing of a
variable displacement vane pump of the third embodiment.
[0034] FIG. 25 is a front elevation view of the variable
displacement vane pump of the third embodiment in the initial
setting state.
[0035] FIG. 26 is a front elevation view of a variable displacement
vane pump of the fourth embodiment in the initial setting
state.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
First Embodiment
[0036] Referring now to the drawings, particularly to FIG. 1, the
variable displacement vane pump (hereinafter is referred to as
"pump VP") of the first embodiment is exemplified in an internal
combustion engine of an automotive vehicle. Pump VP is installed on
the front end of an engine cylinder block, for supplying moving
engine parts with lubricating oil and for delivering oil (serving
as a working medium as well as a lubricating substance) to a
variable valve actuation mechanism, which is installed for variably
controlling operating characteristics of engine valves.
[0037] (Construction of Valve Timing Control System)
[0038] In the first embodiment, a variable valve timing control
(VTC) device is used as the variable valve actuation mechanism (the
phase converter). As seen from the cross section of FIG. 1, in the
shown embodiment, the VTC device is applied to the intake-valve
side of the engine. The VTC system is comprised of a disk-shaped
timing sprocket 100, a camshaft 200, a vane member 300, and a
hydraulic pressure supply-and-exhaust mechanism 400. A phase of
camshaft 200 relative to an engine crankshaft (not shown) can be
continuously varied hydraulically. As shown in FIG. 1, a
hydraulically-operated vane-type timing variator or a
hydraulically-operated vane-type phase converter (simply, a
hydraulic actuator) of the VTC device includes timing sprocket 100,
camshaft 200, and vane member 300. In the cross-sectional view
shown in FIG. 1, assume that the direction of the axis of camshaft
200 is taken as x-axis, one axial direction of camshaft 200
oriented from the left-hand side (viewing FIG. 1) of camshaft 200,
on which timing sprocket 100 is installed, to the right-hand side
of camshaft 200 is defined as a negative x-axis direction, and the
opposite axial direction oriented from the right-hand side (viewing
FIG. 1) to the left-hand side of camshaft 200 is defined as a
positive x-axis direction. FIG. 1 shows the cross section (the
cutaway view) of the hydraulic actuator, cut along the x-axis
direction and the cross section (the cutaway view) of a directional
control valve 450 of hydraulic pressure supply-and-exhaust
mechanism 400, cut along the x-axis direction. FIG. 2 shows the
front elevation of the hydraulic actuator whose front cover is
removed, as viewed from the positive x-axis direction. The cross
section of the hydraulic actuator shown in FIG. 1 corresponds to
the F-G-H cross section taken along the lines F-G and G-H shown in
FIG. 2, but only a portion of an oil seal member 345 is replaced by
the G-I cross section taken along the line G-I rather than the G-H
cross section taken along the line G-H.
[0039] Timing sprocket 100 is driven by the crankshaft via a timing
chain and thus rotates in synchronism with rotation of the
crankshaft. Camshaft 200 is rotatably supported on the upper
portion of an engine cylinder head (not shown) by means of cam
bearings, such that relative rotation of camshaft 200 to timing
sprocket 100 is permitted. Camshaft 200 has a series of cams formed
integral with the camshaft at predetermined axial positions, for
operating (opening and closing) intake valves via respective valve
lifters. Vane member 300 is fixedly connected to the camshaft axial
end (i.e., the left-hand axial end of camshaft 200, viewing FIG.
1), facing in the positive x-axis direction. Vane member 300 is
rotatably accommodated in a phase-converter housing (described
later) of timing sprocket 100. Regarding the hydraulic actuator
(the hydraulically-operated vane-type phase converter), timing
sprocket 100, which rotates in synchronism with rotation of the
crankshaft, serves as a driving rotational member, whereas vane
member 300, which is fixedly connected or bolted to camshaft 200,
serves as a driven rotational member. Hydraulic pressure
supply-and-exhaust mechanism 400 is configured to rotate vane
member 300 in a normal-rotational direction or in a
reverse-rotational direction by hydraulic pressure.
[0040] Timing sprocket 100 has a phase-converter housing 102, a
front cover 103, and a rear cover 104. Housing 102 is formed into a
cylindrical shape, opened at both ends in the opposite x-axis
directions. The outer periphery of housing 102 is formed integral
with a toothed portion 101 in meshed-engagement with the timing
chain. Front cover 103 is installed to hermetically cover the
opening end of housing 102, facing in the positive x-axis
direction, whereas rear cover 104 is installed to hermetically
cover the opening end of housing 102, facing in the negative x-axis
direction. Housing 102, front cover 103, and rear cover 104 are
fastened together with four small-diameter bolts b1-b4.
[0041] Housing 102 is integrally formed on its inner periphery with
four radially-inward protruded shoes 110, 120, 130, and 140. The
four shoes are circumferentially spaced from each other by
approximately 90 degrees. As can be appreciated from the cross
section of FIG. 1, each of shoes 110-140 is formed as an elongated
partition wall portion extending in the x-axis direction of housing
102. These shoes are hereinafter referred to as "partition wall
portions". Each of partition wall portions 110-140 has a
substantially trapezoidal shape in lateral cross section, taken in
the direction perpendicular to the x-axis direction. As seen from
the cross section of FIG. 1, the left-hand axial end face of each
of partition wall portions 110-140 is in wall-contact with the
inner peripheral surface of front cover 103, whereas the right-hand
axial end face of each of partition wall portions 110-140 is in
wall-contact with the inner peripheral surface of rear cover 104.
As seen in FIGS. 1-2, in particular, as clearly seen from the
lateral cross section of each of partition wall portions 110-140 in
FIG. 2, partition wall portions 110-140 are formed substantially at
their centers in trapezoidal lateral cross section with respective
bolt insertion holes 111, 121, 131, and 141 (through holes
extending in the x-axis direction) into which bolts b1, b2, b3, and
b4 are inserted.
[0042] As viewed from the positive x-axis direction, the innermost
ends 112, 122, 132, and 142 of the radially-inward protruded
partition wall portions 110-140 are formed as somewhat concave
circular-arc end faces, which are configured to be substantially
conformable to the shape of the outer periphery of a vane rotor 301
of vane member 300. Partition wall portions 110-140 have respective
axially-elongated seal retaining grooves 113, 123, 133, and 143,
formed in their innermost ends 112, 122, 132, and 142 and extending
in the x-axis direction. Four oil seal members 114, 124, 134, and
144, each being square in lateral cross section, are fitted into
respective seal retaining grooves 113, 123, 133, and 143.
Additionally, four leaf springs (not shown) are retained in
respective seal retaining grooves 113, 123, 133, and 143, in a
manner so as to force four seal members 114, 124, 134, and 144 into
sliding-contact with the outer peripheral surface of vane rotor
301.
[0043] As can be seen from the cross section of FIG. 1, front cover
103 has a centrally-bored, large-diameter bolt insertion hole (a
through hole) 105 into which a cam bolt 211 is inserted.
Additionally, front cover 103 is formed with circumferentially
equidistant-spaced, four bolt holes into which respective bolts
b1-b4 are inserted. As seen from the cross section of FIG. 1, rear
cover 104 is formed at its center with a bearing bore (or a housing
supporting bore) 106, into which the camshaft end 210 of camshaft
200, facing in the positive x-axis direction, is inserted, so that
the inner periphery of rear cover 104 is rotatably supported on the
outer periphery of camshaft end 210. As clearly shown in FIG. 1,
rear cover 104 is formed with circumferentially equidistant-spaced,
four female screw-threaded portions into which the male
screw-threaded portions of bolts b1-b4 are screwed.
[0044] Vane member 300 is rotatably accommodated in the cylindrical
phase-converter housing 102. Vane member 300 is made of metal
materials, such as sintered alloy materials. Vane member 300 is
comprised of a substantially annular ring-shaped vane rotor 301 and
four radially-extending vanes or blades 310, 320, 330, and 340.
Vane rotor 301 and four vane blades 310, 320, 330, and 340 are
integrally formed with each other. Vane rotor 301 has an
axially-extending central bore 302 into which cam bolt (vane
mounting bolt) 211 is inserted for bolting vane member 300 to
camshaft end 210 by axially tightening the cam bolt. The axis of
vane rotor 301 is coaxially aligned with the axis of camshaft 200.
Four blades 310, 320, 330, and 340 are formed integral with vane
rotor 301, such that the four blades are substantially
equidistant-spaced apart from each other in the circumferential
direction of vane rotor 301, and extend radially outwards from the
outer periphery of vane rotor 301. As viewed in the x-axis
direction, vane rotor 300 is formed on its right-hand side, facing
camshaft end 210, with a central cylindrical-hollow fitting groove
303 into which camshaft end 210 is fitted from the negative x-axis
direction.
[0045] As best seen in FIG. 2, in the hydraulically-operated
four-blade vane member equipped VTC device, the areas of the
outside circumferences of four blades 310-340 of the four-blade
vane member 300, in other words, the circumferential widths of four
blades 310-340 are dimensioned or set to be somewhat different from
each other. Four blades 310-340 are classified into two sorts,
namely a maximum-width blade 340 and the remaining narrow-width
blades 310, 320, and 330. The maximum-width blade 340 of the four
vane blades is configured to have an inverted trapezoidal shape in
lateral cross section, whereas the remaining three vane blades 310,
320, and 330 are configured to be substantially rectangular in
lateral cross section. The remaining three blades have almost the
same circumferential width and the same radial length. The
circumferential width of the maximum-width blade 340 having the
inverted trapezoidal shape is dimensioned to be greater than that
of each of the remaining three rectangular blades 310, 320, and
330. The four blades are circumferentially spaced apart from each
other and arranged at predetermined angular positions, taking
account of total weight balance of vane member 300, in other words,
reduced rotational unbalance of vane member 300 having four blades
310-340. Each of four blades 310-340 is disposed in an internal
space defined between the associated two adjacent partition wall
portions.
[0046] Four blades 310-340 have respective axially-elongated seal
retaining grooves 314, 324, 334, and 344, formed in their outermost
ends (apexes) 313, 323, 333, and 343 and extending in the x-axis
direction. Four oil seal members (four apex seals) 315, 325, 335,
and 345 are fitted into respective seal retaining grooves 314, 324,
334, and 344. Additionally, four leaf springs LS (see the cross
section of the hydraulic actuator in FIG. 1) are retained in
respective seal retaining grooves 314, 324, 334, and 344, in a
manner so as to force four seal members 315, 325, 335, and 345 into
sliding-contact with the inner peripheral surface of housing
102.
[0047] The front end face of each of blades 310-340 and the rear
end face of front cover 103 are opposed to each other with a very
small clearance space. In a similar manner, the rear end face of
each of blades 310-340 and the front end face of rear cover 104 are
opposed to each other with a very small clearance space. Four
variable-volume phase-advance chambers 311, 321, 331, and 341 and
four variable-volume phase-retard chambers 312, 322, 332, and 342
are defined among the rear end face of front cover 103, the front
end face of rear cover 104, both sidewalls of each of four blades
310-340 of vane member 300, facing to the rotational direction of
the vane rotor, and both sidewalls of each of four partition wall
portions 110-140 of housing 102. For instance, as seen in FIG. 2,
variable-volume phase-advance chamber 341 is defined between the
sidewall 346 of maximum-width blade 340, facing in the
counterclockwise direction, and the sidewall 145 of partition wall
portion 140, facing in the clockwise direction. Variable-volume
phase-retard chamber 342 is defined between the sidewall 347 of
maximum-width blade 340, facing in the clockwise direction, and the
sidewall 115 of partition wall portion 110, facing in the
counterclockwise direction.
[0048] As shown in FIG. 1, hydraulic pressure supply-and-exhaust
mechanism 400 is comprised of the variable displacement vane pump
VP, directional control valve 450, and a dual hydraulic-line system
410-420. The dual line system is comprised of the first hydraulic
line 410 provided to supply and exhaust working oil (hydraulic
pressure) to and from each of phase-advance chambers 311-341, and
the second hydraulic line 420 provided to supply and exhaust
working oil (hydraulic pressure) to and from each of phase-retard
chambers 312-342. Each of hydraulic lines 410 and 420 are connected
through directional control valve 450 to a working-oil supply
passage 430 (functioning as a main oil gallery) and a working-oil
drain passage 440. Pump VP (serving as a one-way variable
displacement vane pump) is disposed in supply passage 430 for
sucking or drawing working fluid in an oil pan 460 and for
discharging the pressurized working oil from its discharge port.
The downstream end of drain passage 440 communicates oil pan
460.
[0049] First hydraulic line 410 is provided between directional
control valve 450 and each of phase-advance chambers 311-341. First
hydraulic line 410 is further provided with a first flow-passage
structure 411 and a first branch passage structure including four
branch passages 412, 413, 414, and 415. First flow-passage
structure 411 is constructed as a fluid-passage structure extending
from the inside of the cylinder head via the inside of the cam
bearing toward camshaft 200, and partly including an axial oil
passage formed in camshaft 200. Four branch passages 412-415 are
formed in vane rotor 301 in such a manner as to substantially
radially extend from the inner periphery of the cylindrical bore of
vane rotor 301 (see FIG. 2). Four phase-advance chambers 311-341
are communicated with first flow-passage structure 411 via
respective branch passages 412-415.
[0050] On the other hand, second hydraulic line 420 is provided
between directional control valve 450 and each of phase-retard
chambers 312-342. Second hydraulic line 420 is further provided
with a second flow-passage structure 421 and a second branch
passage structure including four branch passages 422, 423, 424, and
425. Second flow-passage structure 421 is constructed as a
fluid-passage structure extending from the inside of the cylinder
head via the inside of the cam bearing toward camshaft 200, and
partly including an axial oil passage formed in camshaft 200. Four
branch passages 422-425 are formed in vane rotor 301 in such a
manner as to substantially radially extend from the inner periphery
of the cylindrical bore of vane rotor 301 (see FIG. 2). Four
phase-retard chambers 312-342 are communicated with second
flow-passage structure 421 via respective branch passages
422-425.
[0051] Directional control valve 450 is constructed by a
spring-offset solenoid-actuated directional control valve.
Directional control valve 450 is comprised of a valve housing (a
substantially cylindrical valve body) 470, an electromagnetic
solenoid 480, and a sliding valve spool 490. Valve housing 470 is
fitted to a valve retaining bore 451 formed in the cylinder head.
Solenoid 480 is installed on the left-hand axial end of valve
housing 470 (viewing FIG. 1). Valve spool 490 is a substantially
cylindrical member that has large-diameter lands (described later)
machined to axially slide in a very close-fitting bore of valve
housing 470. The grooves between the lands provide the flow
passages between ports (described hereunder).
[0052] Valve housing 470 has a supply port 471 formed substantially
at a middle position of the axially-elongated valve housing 470, a
first port 472 formed on the side of the positive x-axis direction
with respect to supply port 471, and a second port 473 formed on
the side of the negative x-axis direction with respect to supply
port 471. Also, valve housing 470 has a first drain port 474 formed
on the side of the negative x-axis direction with respect to first
port 472, and a second drain port 475 formed on the side of the
positive x-axis direction with respect to second port 473. Supply
port 471 intercommunicates supply passage 430 and the internal
space of valve housing 470. First port 472 intercommunicates first
hydraulic line 410 and the internal space of valve housing 470,
whereas second port 473 intercommunicates second hydraulic line 420
and the internal space of valve housing 470. Each of first and
second drain ports 474-475 intercommunicates the internal space of
valve housing 470 and drain passage 440.
[0053] Solenoid 480 is comprised of a solenoid casing 481, an
electromagnetic coil 482 installed in the solenoid casing, a
stationary core 483, and a movable plunger 484. Stationary core 483
is magnetized by energizing electromagnetic coil 482. When
stationary core 483 is magnetized, movable plunger 484 is forced in
the negative x-axis direction against the spring force of a return
spring RS, thus creating a sliding motion of valve spool 490 in the
negative x-axis direction. Electromagnetic coil 482 of solenoid 480
is connected via a wiring harness 485 to an electronic control unit
(simply, a controller) CU.
[0054] Valve spool 490 is formed integral with four lands, that is,
the first land 491, the second land 492 formed on the side of the
negative x-axis direction with respect to first land 491, the third
land 493 formed on the side of the positive x-axis direction with
respect to first land 491, and the fourth land 494 formed on the
side of the positive x-axis direction with respect to third land
493. The left-hand axial end of fourth land 494, facing in the
positive x-axis direction, is in abutted-engagement with the
right-hand axial end of movable plunger 484. Return spring RS is
installed between the right-hand end face of second land 492 and a
return spring retainer 476 formed at the right-hand axial end of
valve housing 470, under preload. As set forth above, directional
control valve 450 uses the sliding valve spool 490 to change the
path of flow through the directional control valve. For a given
position of valve spool 490, a unique flow path configuration
exists within the valve. Concretely, depending on an axial position
of valve spool 490, first land 491 functions to open or close first
port 472, whereas third land 493 functions to open or close second
port 473. More concretely, directional control valve 450 is
designed to switch among at least three positions of the spool,
namely a spring-offset position (a solenoid de-energized position)
shown in FIG. 1, a block-off position (an intermediate position
created due to the balancing opposing forces, that is, the return
spring force and the electromagnetic force produced by the
solenoid), and a fully solenoid-energized position. In the
spring-offset position (with solenoid 480 de-energized), the
maximum displacement of valve spool 490 in the positive x-axis
direction is created by the spring force of return spring RS, and
whereby fluid communication between supply port 471 and second port
473 is established, and fluid communication between first port 472
and first drain port 474 is established. In the block-off position,
valve spool 490 is positioned at its intermediate axial position
between the maximum displacement of valve spool 490 in the positive
x-axis direction and the maximum displacement of valve spool 490 in
the negative x-axis direction, and whereby fluid communication
between each of first and second ports 472-473 and supply port 471
is blocked and fluid communication between each of first and second
ports 472-473 and drain passage 440 is blocked. In the fully
solenoid-energized position, the maximum displacement of valve
spool 490 in the negative x-axis direction is created by the
electromagnetic force of solenoid 480, and whereby fluid
communication between supply port 471 and first port 472 is
established, and fluid communication between second port 473 and
second drain port 475 is established. Switching operation among the
three positions of the valve spool of directional control valve 450
is executed responsively to a control command signal generated from
the output interface circuitry of controller CU to the
solenoid.
[0055] Controller CU generally comprises a microcomputer.
Controller CU includes an input/output interface (I/O), memories
(RAM, ROM), and a microprocessor or a central processing unit
(CPU). The input/output interface (I/O) of the controller receives
input information from various engine/vehicle sensors, namely a
crank angle sensor, an airflow meter, a throttle opening sensor, an
engine temperature sensor (an engine coolant temperature sensor), a
camshaft angular position sensor, and the like. Within the
controller, the central processing unit (CPU) allows the access by
the I/O interface of input informational data signals from the
engine/vehicle sensors. The processor of controller CU determines
the current engine/vehicle operating condition, based on input
information from the engine/vehicle sensors. The crank angle sensor
is provided to detect an angular position (crankangle) of the
crankshaft, and for detecting engine speed. The camshaft angular
position sensor is provided for detecting an angular position of
camshaft 200. Also, based on both of the sensor signals from the
crank angle sensor and the camshaft angular position sensor, an
angular phase of camshaft 200 relative to timing sprocket 100 is
detected. The airflow meter is provided for measuring or detecting
a quantity of air flowing through an intake pipe, and consequently
for detecting or estimating the magnitude of engine load. The CPU
of the controller is responsible for carrying the phase control
program stored in memories. Computational results (arithmetic
calculation results), that is, a calculated output signal (e.g., a
pulsed control current) is relayed through the output interface
circuitry of the controller to output stages, namely the solenoid
480 (exactly, the electrically energized solenoid coil 482) of
electromagnetic directional control valve 450.
[0056] Also provided is a lock mechanism 500 disposed between
maximum-width blade 340 of vane member 300 and rear cover 104 of
phase-converter housing 102, for disabling rotary motion of vane
member 300 relative to rear cover 104 (or timing sprocket 100) by
locking and engaging vane member 300 with housing 102, and for
enabling rotary motion of vane member 300 relative to rear cover
104 by unlocking (or disengaging) vane member 300 from housing 102.
As can be seen from the cross section of FIG. 1, lock mechanism 500
is comprised of a lock piston 510, an engaging-hole structural
member 520, a spring retainer 530, and a return spring (a coiled
compression spring) 540. FIG. 14B shows the cross section, taken
along the line J-J of FIG. 2.
[0057] A lock-piston sliding-motion permitting bore (simply, a
lock-piston bore) 501 is formed in the inverted trapezoidal blade
340 of the maximum circumferential width, such that lock-piston
bore 501 extends in the x-axis direction of camshaft 200.
Lock-piston bore 501 is comprised of a small-diameter chamber 502
formed on the side of the negative x-axis direction and a
large-diameter chamber 503 formed on the side of the positive
x-axis direction. Lock piston 510 is formed into a substantially
cylindrical shape and closed at one axial end (the right-hand side
axial end, viewing in FIG. 1). Lock piston 510 is slidably
installed in lock-piston bore 501. The right-hand axial end of lock
piston 510 is formed as a tapered head portion 511, which is
engaged with or disengaged from engaging-hole structural member 520
of rear cover 104. Lock piston 510 has a cylindrical sliding
portion 512 formed integral with tapered head portion 511 and
extending from tapered head portion 511 in the positive x-axis
direction. The leftmost end of lock piston 510 is formed as an
annular flanged portion 513.
[0058] The outside diameter of sliding portion 512 is dimensioned
to be substantially equal to the inside diameter of small-diameter
chamber 502 of lock-piston bore 501. Sliding portion 512 is
accommodated in small-diameter chamber 502, such that sliding
motion of sliding portion 512 relative to small-diameter chamber
502 is permitted. The outside diameter of annular flanged portion
513 is dimensioned to be greater than the outside diameter of
sliding portion 512 and also dimensioned to be substantially equal
to the inside diameter of large-diameter chamber 503 of lock-piston
bore 501. Flanged portion 513 is accommodated in large-diameter
chamber 503, such that sliding motion of flanged portion 513
relative to large-diameter chamber 503 is permitted. A stepped
portion 504 is formed between small-diameter chamber 502 and
large-diameter chamber 503 of maximum-width blade 340. A
pressure-receiving chamber 550 is defined between the annular face
of stepped portion 504, facing in the positive x-axis direction,
and the annular face of flanged portion 513, facing in the negative
x-axis direction.
[0059] Rear cover 104 is formed with an axially-bored retaining
hole 505. Engaging-hole structural member 520 has a cup-shape in
axial cross section, and press-fitted into the retaining hole 505
of rear cover 104. A lock-piston engaging hole 521, having a
substantially trapezoidal shape in axial cross section, is defined
in the cup-shaped engaging-hole structural member 520. As seen in
FIG. 2, when vane member 300 is positioned in its maximum
phase-retard position, that is, when the sidewall 346 of
maximum-width blade 340 is kept in abutted-engagement with the
sidewall 145 of partition wall portion 140 and thus the volumetric
capacity of phase-advance chamber 341 becomes minimum, as viewed
from the x-axis direction, the axis of lock piston 510 is aligned
with the axis of engaging hole 521 of cup-shaped engaging-hole
structural member 520 of rear cover 104. Under these conditions,
when a sliding motion of lock piston 510 in the negative x-axis
direction occurs, tapered head portion 511 of lock piston 510 is
brought into engagement with engaging hole 521. Conversely when a
sliding motion of lock piston 510 in the positive x-axis direction
occurs, tapered head portion 511 of lock piston 510 is brought out
of engagement with engaging hole 521.
[0060] Spring retainer 530 is fitted to the inner peripheral
surface of large-diameter chamber 503. Return spring 540 is
installed between spring retainer 530 and lock piston 510, under
preload. Return spring 540 acts to permanently force lock piston
510 in the negative x-axis direction, that is, toward rear cover
104 (i.e., toward lock-piston engaging hole 521). With vane member
300 kept in its maximum phase-retard position (see FIG. 2), tapered
head portion 511 of lock piston 510 is forced in the negative
x-axis direction by the spring force of return spring 540, and thus
tapered head portion 511 is brought into engagement with engaging
hole 521. As a result of this, relative rotation of vane member 300
to rear cover 104, that is, relative rotation of camshaft 200 to
timing sprocket 100, is locked or prevented.
[0061] As shown in FIG. 14B, a first oil groove 343g is formed in
maximum-width blade 340, so as to intercommunicate phase-advance
chamber 341 and small-diameter chamber 502. A second oil hole 344h
is also formed in maximum-width blade 340, so as to
intercommunicate phase-retard chamber 342 and pressure-receiving
chamber 550. Tapered head portion 511 of lock piston 510 receives
hydraulic pressure of working oil supplied from phase-advance
chamber 341 via first oil groove 343g into lock-piston engaging
hole 521, such that lock piston 510 is forced in the positive
x-axis direction by the hydraulic pressure. Additionally, flanged
portion 513 of lock piston 510 receives hydraulic pressure of
working oil from phase-retard chamber 342 via second oil hole 344h
into pressure-receiving chamber 550, such that lock piston 510 is
forced in the positive x-axis direction by the hydraulic pressure.
Hence, lock piston 510 can be forced in the positive x-axis
direction by either the hydraulic pressure applied to tapered head
portion 511 or the hydraulic pressure applied to flanged portion
513, against the spring force of return spring 540, so as to
disengage lock piston 510 from lock-piston engaging hole 521.
[0062] As discussed above, return spring 540 functions as a
locked-state holding mechanism. The spring force (i.e., the spring
stiffness) of return spring 540 is designed or set such that lock
piston 510 cannot be disengaged from lock-piston engaging hole 521
without a remarkable compressive deformation of return spring 540,
even when air staying in phase-retard chamber 342 during an engine
startup period is compressed by hydraulic pressure of working oil
force-fed from pump VP to phase-retard chamber 342, and then the
compressed air is introduced into pressure-receiving chamber 550 to
force flanged portion 513 of lock piston 510 in the positive x-axis
direction.
[0063] (Construction of Variable Displacement Vane Pump)
[0064] As seen from the disassembled view of pump VP of FIG. 3,
pump VP is comprised of a pump housing 1, a pump cover 2, a drive
shaft 3, a pump rotor 4, a cam ring 5, a plurality of vanes 6, a
pair of vane rings 7 (i.e., vane rings 7a, 7b), a biasing member 8,
and a pivot pin 9. For convenience's sake, assume that the
direction of the axis "O" of drive shaft 3 is taken as z-axis, one
axial direction of drive shaft 3 oriented from pump housing 1 to
pump cover 2 is defined as a positive z-axis direction, and the
opposite axial direction of drive shaft 3 oriented from pump cover
2 to pump housing 1 is defined as a negative z-axis direction.
[0065] Pump housing 1 is formed into a substantially cylindrical
shape and closed at one axial end (the right-hand side axial end,
viewing in FIG. 3). Under a condition where vane-pump component
parts, such as rotor 4 and cam ring 5, are accommodated or
installed in pump housing 1, the left-hand side opening end of pump
housing 1 is hermetically closed by pump cover 2. Pump housing 1
has a basal portion 10 formed on the side of the negative z-axis
direction, a peripheral wall 13 extending from the perimeter of
basal portion 10 in the positive z-axis direction, and a flanged
portion 14 formed on the side of the positive z-axis direction of
peripheral wall 13. Basal portion 10, peripheral wall 13, and
flanged portion 14 are formed integral with each other. Basal
portion 10 is formed at a substantially central portion with a
bearing bore (or a drive-shaft supporting bore) 11, by which drive
shaft 3 is rotatably supported. Bearing bore 11 is formed as a
through hole penetrating basal portion 10 in the z-axis direction.
Additionally, basal portion 10 has a pin insertion hole 12 whose
axis is parallel to the z-axis. Pivot pin 9 is inserted and fitted
into pin insertion hole 12 from the positive z-axis direction. The
perimeter of flanged portion 14 is formed with seven female
screw-threaded portions 14a-14g into which the male screw-threaded
portions of bolts B1-B7 are screwed.
[0066] As viewed from the z-axis direction, pump cover 2 has almost
the same shape as pump housing 1. Pump cover 2 is comprised of a
main-body portion 20, and a flanged portion 24 formed integral with
the perimeter of main-body portion 20. Main-body portion 20 is
formed at a substantially central portion with a bearing bore (or a
drive-shaft supporting bore) 21, by which drive shaft 3 is
rotatably supported, and which is formed as a through hole
penetrating main-body portion 20 in the z-axis direction.
Additionally, the bottom face (or the base) 20a of main-body
portion 20 has a pin supporting portion 20b formed on the side of
the negative z-axis direction, so as to support the axial end of
pivot pin 9, facing in the positive z-axis direction. As described
later, portions formed in bottom face 20a of main-body portion 20
of pump cover 2 and indicated by the broken line in FIG. 3,
correspond to an inlet port 22 (described later), a discharge port
23 (described later), and oil storage portions (described
later).
[0067] The perimeter of flanged portion 24 is formed with seven
bolt holes, which are formed as through holes penetrating the
perimeter of flanged portion 24 in the z-axis direction. Seven
bolts B1-B7 are inserted into the respective bolt holes of flanged
portion 24 of pump cover 2 and then the male screw-threaded
portions of bolts B1-B7 are screwed into respective female
screw-threaded portions 14a-14g of pump housing 1, such that pump
cover 2 is fixedly connected to pump housing 1 by tightening seven
bolts B1-B7. For the reasons discussed later, in the shown
embodiment, pump housing 1 and pump cover 2 are integrally
connected to each other with seven bolts B1-B7 without interleaving
a seal member (an oil seal or a gasket usually used to enhance a
fluid-tight performance of the vane pump) between flanged portion
14 of pump housing 1 and flanged portion 24 of pump cover 2.
[0068] Both ends of drive shaft 3 are inserted into respective
bearing bores 11 and 21 of pump housing 1 and pump cover 2, such
that drive shaft 3 is rotatably supported by means of bearing bores
11 and 21. Rotor 4 is fixed onto the outer periphery of drive shaft
3 for co-rotation with drive shaft 3. The axial end of drive shaft
3, facing in the negative z-axis direction, is connected to the
engine crankshaft. That is, drive shaft 3 of pump VP is driven by
torque transmitted from the engine crankshaft to drive shaft 3. As
viewed from the positive z-axis direction, drive shaft 3 rotates
counterclockwise.
[0069] Rotor 4 has a substantially cylindrical shape or a
substantially disc shape. Assuming that rotor 4 is cut along the
plane passing through the axis of rotor 4, the cross section
becomes a substantially I-shaped cross section. That is, rotor 4
has a thin-walled inner peripheral portion 41 having a
comparatively less thickness in the z-axis direction, and a
thick-walled outer peripheral portion 42 having a comparatively
greater thickness in the z-axis direction. The thin-walled inner
peripheral portion 41 of rotor 4 is formed with a fitted central
bore 40 (a through hole) penetrating the center of rotor 4 in the
z-axis direction. Drive shaft 3 and rotor 4 are integrally
connected to each other by press-fitting drive shaft 3 into the
fitted central bore 40 of rotor 4. Rotor 4 is rotatably
accommodated in pump housing 1. Rotor 4, together with drive shaft
3, is driven by the engine crankshaft. That is, rotor 4, together
with drive shaft 3, rotates in synchronism with rotation of the
crankshaft.
[0070] As best seen in FIG. 3, rotor 4 is formed with
circumferentially equidistant-spaced, radially-extending seven
slits 4a-4g. Seven slits 4a-4g, each having a predetermined
circumferential width in the circumferential direction of rotor 4,
are formed in rotor 4, in such a manner as to extend radially
inwards from the outer peripheral surface 42a of thick-walled outer
peripheral portion 42 toward the axis "O" of drive shaft 3 without
reaching the fitted central bore 40. Radially-inward extending
slits 4a-4g are formed at their innermost ends with respective
back-pressure chambers 40a-40g, each having a substantially
circular in lateral cross section.
[0071] Cam ring 5 is a movable member, which is installed in a
manner so as to be slidable relative to each of pump cover 1 and
pump housing 2, while accommodating therein rotor 4. Cam ring 5 is
substantially cylindrical in shape. Cam ring 5 is formed of a
cylindrical portion 5a, a sector portion 5b, a pivot portion 5c,
and an arm portion 5d. These portions 5a-5d are formed integral
with each other and made of sintered alloy materials, such as
iron-based sintered alloy materials.
[0072] Cylindrical portion 5a accommodates therein rotor 4. As
viewed from the z-axis direction, assume that the geometric center
of an inner peripheral surface (a cylinder bore) 50 of cylindrical
portion 5a is taken as a point "P". Sector portion 5b is laid out
on the outer periphery of cylindrical portion 5a and formed
integral with cylindrical portion 5a. Sector portion 5b has a
substantially sector cross-section in the direction along the plane
perpendicular to the z-axis. Sector portion 5b is formed therein
with a working-oil communication hole 51.
[0073] In a similar manner to sector portion 5b, pivot portion 5c
is also laid out on the outer periphery of cylindrical portion 5a
and formed integral with cylindrical portion 5a. Pivot portion 5c
has a small annular cross section in the direction along the plane
perpendicular to the z-axis. Pivot portion 5c has a pivot bore 52
formed as a through hole extending in the z-axis direction. Cam
ring 5 is accommodated in the internal space of pump housing 1,
under a condition where pivot pin 9 is inserted and fitted into
pivot bore 52. Cam ring 5 is rotatably supported by means of pivot
portion 5c in such a manner as to be rotatable about the pivot pin
9. That is, pivot pin 9 serves as a pivot of cam ring 5, in other
words, a fulcrum of oscillating motion of cam ring 5.
[0074] Arm portion 5d and pivot portion 5c are laid out to be
substantially symmetrical to each other with respect to the
geometric center "P". Arm portion 5d is also laid out on the outer
periphery of cylindrical portion 5a and formed integral with
cylindrical portion 5a. The width of cam ring 5 in the z-axis
direction is the same for all of cylindrical portion 5a, sector
portion 5b, pivot portion 5c, and arm portion 5d. The width of cam
ring 5 in the z-axis direction is dimensioned to be substantially
identical to the depth of pump housing 1, that is, the length of
peripheral wall 13 of pump housing 1 in the z-axis direction.
[0075] The end face of cam ring 5, facing in the positive z-axis
direction, and the end face of pump-cover main-body portion 20,
facing in the negative z-axis direction, are opposed to each other
with a very small clearance space. Thus, sliding motion of the end
face of cam ring 5, facing in the positive z-axis direction,
relative to the end face of pump-cover main-body portion 20, facing
in the negative z-axis direction, is permitted. In a similar
manner, the end face of cam ring 5, facing in the negative z-axis
direction, and the bottom face 10a of pump-housing basal portion
10, facing in the positive z-axis direction, are opposed to each
other with a very small clearance space. Thus, sliding motion of
the end face of cam ring 5, facing in the negative z-axis
direction, relative to the bottom face 10a of pump-housing basal
portion 10, facing in the positive z-axis direction, is permitted.
That is, pump-housing basal portion 10 and pump-cover main-body
portion 20 are assembled or installed to serve as sidewalls
opposing both sides of cam ring 5 in the opposite z-axis
directions. Cam ring 5 is provided between the two opposed
sidewalls of pump-housing basal portion 10 and pump-cover main-body
portion 20, such that oscillating motion of cam ring 5 about the
pivot (i.e., pivot pin 9) is permitted.
[0076] In the shown embodiment, the term "sliding motion" basically
means that two members are in sliding-contact with each other, such
that a relative displacement of one of the two members to the other
is permitted. The term "sliding motion" also means that two members
are in sliding-contact with each other via an oil film filling in
the clearance space defined between them, such that a relative
displacement of one of the two members to the other is permitted
with the oil film for lubrication.
[0077] Cam ring 5 is configured so that the geometric center "P" of
the cylinder bore of cam-ring cylindrical portion 5a can displace
from the axis "O" of drive shaft 3 in the direction perpendicular
to the axis "O" of drive shaft 3, while keeping a parallel layout
of the geometric center "P" of the cylinder bore (inner peripheral
surface 50) of cam-ring cylindrical portion 5a parallel to the axis
"O" of drive shaft 3 in the z-axis direction. That is, cam ring 5
is configured so that the geometric center of "P" of cam ring 5 can
oscillate eccentrically with respect to the axis "O" of drive shaft
3.
[0078] In the shown embodiment, the plurality of vanes 6 of pump VP
are seven vanes 6a, 6b, 6c, 6d, 6e, 6f, and 6g. These vanes 6a-6g
are the same in shape and formed into a rectangular shape. The
width of each of vanes 6a-6g is dimensioned to be substantially
identical to the length of rotor 4 in the z-axis direction. Vane 6a
is fitted into the associated slit 4a of rotor 4, in such a manner
as to be slidable (retractable and extendable) in the radial
direction of rotor 4. In the same manner as vane 6a, the other
vanes 6b-6g are slidably fitted into respective slits 4b-4g. The
length of each of vanes 6a-6g in the radial direction of rotor 4 is
dimensioned to be shorter than the overall depth of each of slits
4a-4g, including respective back-pressure chambers 40a-40g. Each of
vanes 6a-6g is slidably fitted into respective slits 4a-4g, so as
to be radially extendable from outer peripheral surface 42a of
rotor 4 toward inner peripheral surface 50 of cylindrical portion
5a.
[0079] The vane-ring pair 7 is comprised of two ring-shaped members
7a, 7b, each having the same shape and has the same outside
diameter dimensioned to be smaller than the outside diameter of
inner peripheral portion 41 of rotor 4. Vane ring 7a is installed
in one sidewall of inner peripheral portion 41 from the positive
z-axis direction, so that sliding motion of vane ring 7a relative
to the one sidewall of inner peripheral portion 41 is permitted.
Vane ring 7b is installed in the opposite sidewall of inner
peripheral portion 41, so that sliding motion of vane ring 7b
relative to the opposite sidewall of inner peripheral portion 41 is
permitted. Drive shaft 3 extends in the z-axis direction so as to
pass through the internal space of each of vane rings 7a, 7b. The
radially-inward end (the root) of each of vanes 6a-6g is in
abutted-engagement with each of the outer peripheral surfaces of
vane rings 7a-7b.
[0080] By means of the abutted portions of vane rings 7a-7b, each
of vanes 6a-6g is supported with two points in the z-axis
direction. The vane-ring pair 7a-7b has a function that pushes or
forces each of vanes 6a-6g outwards in the radial direction of
rotor 4. The tip (the top end) of each of the radially-outward
forced vanes 6a-6g is in abutted-engagement with inner peripheral
surface 50 of cylindrical portion 5a.
[0081] That is, when the center of each of vane rings 7a-7b is
aligned with the geometric center "P" of cam ring 5, the distance
between inner peripheral surface 50 of cylindrical portion 5a and
each of outer peripheral surfaces 70a-70b of vane rings 7a-7b is
dimensioned to be substantially identical to the length of each of
vanes 6a-6g in the radial direction of rotor 4. Therefore, pump VP
is configured such that, during rotation of rotor 4, the root of
each of vanes 6a-6g is kept in sliding-contact with the outer
peripheral surfaces 70a-70b of vane rings 7a-7b, while the tip of
each of vanes 6a-6g is kept in sliding-contact with inner
peripheral surface 50 of cam ring 5. In other words, during
rotation of rotor 4, the center of each of vane rings 7a-7b is
automatically positioned so as to align with the geometric center
"P" of cam ring 5 by abutment of the root of each of vanes 6a-6g
with the outer peripheral surfaces 70a-70b of vane rings 7a-7b.
[0082] Biasing member 8 is comprised of a small-diameter first coil
spring 8a and a large-diameter second coil spring 8b. Biasing
member 8 is accommodated in a spring chamber 15d defined in pump
housing 1, under preload. Biasing member 8 forces arm portion 5d of
cam ring 5 in one direction by a biasing force (a spring bias or a
spring force), so as to produce a moment by which cam ring 5 can be
rotated about pivot pin 9. Biasing member 8 is installed in
pump-housing spring chamber 15d so as to permanently force cam ring
5 in the one direction (in the direction of action of spring bias)
in which the eccentricity of cam ring 5 increases, in other words,
the geometric center "P" of the cylinder bore of cam-ring
cylindrical portion 5a displaces apart from the axis "O" of drive
shaft 3 (i.e., the rotation center "O" of vane rotor 4).
[0083] (Construction of Pump Housing)
[0084] FIG. 4 shows the front elevation view of pump housing 1, as
viewed from the positive z-axis direction of FIG. 3. In the
elevation view of FIG. 4, assume that the center of bearing bore 11
of basal portion 10 of pump housing 1 (i.e., the axis of drive
shaft 3) is taken as "origin O", a directed line Ox is taken as
x-axis, and a directed line Oy is taken as y-axis to set an
orthogonal coordinate system. In FIG. 4, the line parallel to the
bottom face of the inner periphery of a second swelling portion 1c
(i.e., the bottom face 15e of spring chamber 15d) and passing
through the origin "O" (the axis of drive shaft 3 or the rotation
center of rotor 4) is taken as the x-axis, whereas the line
perpendicular to bottom face 15e and passing through the origin "O"
is taken as the y-axis. The side of spring chamber 15d with respect
to the origin "O" is defined as a positive x-axis direction. The
side of a discharge hole 17a with respect to the origin "O" (or the
x-axis) is defined as a positive y-axis direction.
[0085] Basal portion 10, peripheral wall 13, and flanged portion
14, constructing pump housing 1, are formed integral with each
other, and made of aluminum alloy materials. During oscillating
motion of cam ring 5, the end face of cam ring 5, facing in the
negative z-axis direction, slides along the bottom face 10a of
pump-housing basal portion 10, facing in the positive z-axis
direction. Thus, the area of the bottom face 10a of pump-housing
basal portion 10, corresponding to a given area of sliding motion
of cam ring 5, is more accurately machined in flatness and surface
roughness.
[0086] Pump housing 1 has a cylindrical portion 1a, and first and
second swelling portions 1b-1c. As viewed from the z-axis
direction, the inner peripheral surface 13a of peripheral wall 13
of cylindrical portion 1a is formed into a substantially circular
shape whose center is the origin "O". The distance, measured from
the origin "O" to inner peripheral surface 13a in the negative
y-axis direction, is dimensioned to be slightly greater than the
distance, measured from the origin "O" to inner peripheral surface
13a in the positive y-axis direction. Pump-housing cylindrical
portion 1a is configured to accommodate therein cam-ring
cylindrical portion 5a. First swelling portion 1b is formed to
swell radially outwards from pump-housing cylindrical portion 1a in
a combined direction of the negative x-axis direction and the
positive y-axis direction. In other words, first swelling portion
1b is laid out within the second quadrant of the orthogonal
coordinate system, which second quadrant is defined as {(x,
y)|x<0, y>0}. Sector portion 5b and pivot portion 5c of cam
ring 5 are both accommodated in first swelling portion 1b.
[0087] Second swelling portion 1c is formed to swell radially
outwards from pump-housing cylindrical portion 1a in the positive
x-axis direction. Second swelling portion 1c is formed as a hollow
rectangular parallelopiped. Second swelling portion 1c has an
arm-portion accommodating chamber 15a formed on the side of the
positive y-axis direction of second swelling portion 1c, and spring
chamber 15d formed on the side of the negative y-axis direction of
second swelling portion 1c. Arm-portion accommodating chamber 15a
accommodates therein arm portion 5d of cam ring 5, whereas spring
chamber 15d accommodates therein biasing member 8.
[0088] As viewed from the z-axis direction, the inner peripheral
surface of arm-portion accommodating chamber 15a is formed into a
substantially rectangular shape. Arm-portion accommodating chamber
15a is configured to be surrounded by a seat surface 15b arranged
parallel to the x-axis on the side of the positive y-axis direction
of arm-portion accommodating chamber 15a and a wall surface 15c
parallel to the y-axis on the side of the positive x-axis direction
of arm-portion accommodating chamber 15a. Arm-portion accommodating
chamber 15a is configured to open into spring chamber 15d in the
negative y-axis direction. Seat surface 15b is formed at the
position substantially symmetrical to the pin insertion hole 12
with respect to the origin "O". Concretely, seat surface 15b is
formed substantially at the same level as the center of pin
insertion hole 12 in the y-axis direction. In an initial setting
state of cam ring 5 installed in pump housing 1, seat surface 15b
functions as a seat on which arm portion 5d is seated. In order to
suppress initial fluctuations in a pump discharge of pump VP, seat
surface 15b is more accurately machined, fully taking into account
the positional relationship with both pin insertion hole 12 and
bearing bore 11.
[0089] As viewed from the z-axis direction, the inner peripheral
surface of spring chamber 15d is formed into a substantially
rectangular recessed shape. Spring chamber 15d is configured to be
surrounded in three directions by two wall surfaces 15f-15g, both
parallel to the y-axis, and bottom face 15e parallel to the x-axis.
Spring chamber 15d is configured to open into arm-portion
accommodating chamber 15a on the side of the positive y-axis
direction of spring chamber 15d. Two shoulder portions (engaging
portions) 15h-15i, extending in the z-axis direction and opposed to
each other in the x-axis direction, are formed at the opening end
of spring chamber 15d. Shoulder portion 15h, located on the side of
the negative x-axis direction of spring chamber 15d, is formed to
protrude by a predetermined length in the positive x-axis direction
from the uppermost end (viewing FIG. 4) of a peripheral wall
portion 13b by which wall surface 15f is defined. Shoulder portion
15i, located on the side of the positive x-axis direction of spring
chamber 15d, is formed to protrude by a predetermined length in the
negative x-axis direction from the rightmost end (viewing FIG. 4)
of peripheral wall 13 by which wall surface 15g is defined.
[0090] Pump housing 1 has an inlet portion (namely, an inlet hole
16a, an inlet port 16b), a discharge portion (namely, a discharge
hole 17a, a discharge port 17b), oil storage portions 18a-18c,
which are collectively referred to as "oil storage portion 18", and
a bearing lubrication oil groove 18d, all formed in pump-housing
basal portion 10, in addition to bearing bore 11 and pin insertion
hole 12.
[0091] Inlet hole 16a is formed as a cylindrical through opening,
which penetrates basal portion 10 in the z-axis direction. Inlet
hole 16a is located on the side of the positive x-axis direction of
cylindrical portion 1a in such a manner as to be slightly offset
from the directed line Ox in the negative y-axis direction. Inlet
hole 16a is arranged to bestride the boundary between the rightmost
end of cylindrical portion 1a and the leftmost end of second
swelling portion 1c. As viewed from the z-axis direction, inlet
hole 16a is configured to overlap with a part of peripheral wall
portion 13b and shoulder portion 15h. Inlet hole 16a serves as a
working-oil inlet passage when drawing working oil stored in oil
pan 460 into the pump during operation of pump VP.
[0092] Inlet port 16b is a crescent-shaped groove formed in
pump-housing basal portion 10 and having a predetermined depth and
a predetermined width. Inlet port 16b is arranged on the right-hand
half of bottom face 10a (on the side of the positive x-axis
direction of cylindrical portion 1a). As viewed from the z-axis
direction, inlet port 16b is formed in bottom face 10a as a
circular arc with the center "O" (corresponding to the axis of
drive shaft 3) and a predetermined distance (i.e., a predetermined
radius) from the center "O". The circular-arc shaped inlet port 16b
is arranged to be symmetrical with respect to the x-axis so as to
extend circumferentially by approximately 120 degrees. Inlet port
16b is formed on the side of the negative x-axis direction of seat
surface 15b (or arm-portion accommodating chamber 15a). Inlet port
16b communicates inlet hole 16a.
[0093] Discharge hole 17a is formed in basal portion 10 as a
cylindrical opening, which extends in the z-axis direction.
Discharge hole 17a is located within first swelling portion 1b.
Discharge hole 17a serves as a working-oil discharge passage when
discharging working oil from pump VP during operation of pump VP.
Discharge hole 17a is connected to supply passage 430 (functioning
as the main oil gallery of the engine), so as to communicate with
moving and/or sliding engine parts and the VTC device.
[0094] Discharge port 17b is a groove formed in pump-housing basal
portion 10 and having a predetermined depth. Discharge port 17b is
comprised of a circular-arc shaped groove 17c arranged on the
left-hand half of bottom face 10a (on the side of the negative
x-axis direction of cylindrical portion 1a) and having a
predetermined circumferential width, and a sector groove 17d formed
in bottom face 10a of first swelling portion 1b in such a manner as
to be continuous with circular-arc shaped groove 17c. As viewed
from the z-axis direction, sector groove 17d is configured to
overlap with discharge hole 17a. Discharge port 17b communicates
discharge hole 17a. When viewed from the z-axis direction,
discharge hole 17a is formed to open into only the sector groove
17d. Discharge hole 17a communicates with the interior space of
pump housing 1 via discharge port 17b (sector groove 17d).
[0095] Circular-arc shaped groove 17c of discharge port 17b is
arranged to be symmetrical to the crescent-shaped inlet port 16b
with respect to the center "O", and having almost the same shape as
inlet port 16b, and arranged on the side of the positive x-axis
direction with respect to pin insertion hole 12. Sector portion 17d
is configured to be surrounded in three directions by a side 17g
parallel to the y-axis and located on the side of the negative
x-axis direction of sector portion 17d, a large circular arc 17e
whose center is the geometric center "Q" of pin insertion hole 12
and which is located on the side of the positive y-axis direction
of sector portion 17d, and a small circular arc 17e whose center is
the geometric center "Q" of pin insertion hole 12 and which is
located on the side of the negative y-axis direction of sector
portion 17d in such a manner as to be opposed to large circular arc
17e. Sector groove 17d is opened in the positive x-axis direction
so as to communicate circular-arc shaped groove 17c.
[0096] A substantially cylindrical support portion 12a is formed on
the side of the negative y-axis direction of first swelling portion
1b. Pin insertion hole 12 is formed in support portion 12a. The
geometric center "Q" of pin insertion hole 12 is located to be
slightly offset from the x-axis in the positive y-axis direction by
a predetermined distance. The outer periphery of support portion
12a, facing in the positive y-axis direction, is contoured to
define the previously-noted small circular arc 17f. The outer
periphery of support portion 12a, facing in the positive x-axis
direction, is contoured to define a left-hand curved portion of
circular-arc shaped groove 17c (a circumferentially-curved outside
edged portion formed on the side of the negative x-axis direction
of circular-arc shaped groove 17c).
[0097] The circumferential end "A" of circular-arc shaped inlet
port 16b, facing in the counterclockwise direction and the
circumferential end "D" of circular-arc shaped groove 17c of
discharge port 17b, facing in the counterclockwise direction are
point-symmetrical with respect to the center "O". In a similar
manner, the circumferential end "B" of circular-arc shaped inlet
port 16b, facing in the clockwise direction and the circumferential
end "C" of circular-arc shaped groove 17c of discharge port 17b,
facing in the clockwise direction are point-symmetrical with
respect to the center "O". Therefore, the angle .angle.AOC is
nearly equal to the angle .angle.BOD, that is,
.angle.AOC.apprxeq..angle.BOD. As a result of the tuned positions
of two ports 16b and 17b, these two angles .angle.AOC and
.angle.BOD may not be analogous to each other. As previously
discussed, the geometric center "Q" of pin insertion hole 12 is
located to be slightly offset from the x-axis in the positive
y-axis direction by a predetermined distance, and thus the angle
.angle.DOQ is dimensioned to be greater than the angle .angle.COQ,
i.e., .angle.DOQ>.angle.COQ.
[0098] Oil storage portion 18 is a substantially crescent-shaped
groove formed in pump-housing basal portion 10 and having a
predetermined depth and a predetermined width. Oil storage portion
18 is comprised of three oil storage portions 18a-18c. Oil storage
portions 18a-18c formed in bottom face 10a of cylindrical portion
1a and arranged on the outer peripheral side of bearing bore 11 and
on the inner peripheral side of each of inlet port 16b and
discharge port 17b. The center common to oil storage portions
18a-18c is the center (or the origin) "O". Three oil storage
portions 18a-18c are arranged to be circumferentially
equidistant-spaced from each other so as to surround the center
"O", i.e., the circumference of bearing bore 11. Oil storage
portions 18a and 18c are laid out to be symmetrical to each other
with respect to the x-axis, so as to be partly opposed to discharge
port 17b. Oil storage portion 18b is laid out to be opposed to
inlet port 16b. With rotor 4 installed in pump housing 1, as viewed
from the z-axis direction, oil storage portions 18a-18c are
configured to overlap with inner peripheral portion 41 of rotor
4.
[0099] Oil storage portions 18a-18c temporarily store working oil
discharged from discharge port 17b, and deliver working oil via
bearing lubrication oil groove 18d to bearing bore 11, and also
deliver working oil to the sidewalls of rotor 4, facing apart from
each other in the z-axis direction and to the sidewalls of each of
vanes 6, facing apart from each other in the z-axis direction. This
contributes to the enhanced lubricating performance of pump VP.
[0100] Bearing lubrication oil groove 18d is an oil supply groove
formed in pump-housing basal portion 10 and having a predetermined
depth. Bearing lubrication oil groove 18d is formed in bottom face
10a of cylindrical portion 1a in a manner so as to extend
substantially midway between two oil storage portions 18a and 18c.
Bearing lubrication oil groove 18d intercommunicates discharge port
17b and bearing bore 11. Concretely, bearing lubrication oil groove
18d is formed into a substantially doglegged shape (as viewed from
the z-axis direction) in order to prevent the radially-slidable
vane 6, rotating about the axis "O" (the rotation center of rotor
4), from dropping into bearing lubrication oil groove 18d when the
radially-extending portion of bearing lubrication oil groove 18d
becomes aligned with the radially-slidable vane 6, rotating about
the axis "O". Bearing lubrication oil groove 18d is comprised of an
oblique oil passage extending from discharge port 17b and oriented
in a combined direction of the positive x-axis direction and the
negative y-axis direction, and a radial oil passage extending in
the positive x-axis direction from the substantially midpoint of
two oil storage portions 18a and 18c, which are symmetrical with
respect to the x-axis, and reaching bearing bore 11. Bearing
lubrication oil groove 18d functions to feed working oil from each
of discharge port 17b and oil storage portions 18a and 18c to
bearing bore 11, thus ensuring the lubricating performance of drive
shaft 3.
[0101] In the same manner as pump housing 1, main-body portion 20
and flanged portion 24, both constructing pump cover 2, are formed
integral with each other, and made of aluminum alloy materials. As
indicated by the broken line in FIG. 3, pump cover 2 has inlet port
22, discharge port 23 (sector groove 23d), and oil storage
portions, all formed in bottom face 20a of pump-cover main-body
portion 20, and configured to be substantially conformable to the
respective shapes (the respective sizes and positions) of inlet
port 16b, discharge port 17b (sector groove 17d), and oil storage
portions 18a-18c, formed in bottom face 10a of pump-housing basal
portion 10. Bottom face 20a, in sliding-contact with the end face
of cam ring 5, facing in the positive z-axis direction, is more
accurately machined in flatness and surface roughness.
[0102] (Construction of Pump Chamber)
[0103] FIG. 5 shows the front elevation view of pump VP whose pump
cover is removed, as viewed from the positive z-axis direction. The
positions of inlet hole 16a, inlet port 16b, discharge hole 17a,
and discharge port 17b are indicated by the broken line in FIG. 5.
FIG. 5 shows the initial setting state of cam ring 5, where the
oscillated amount of cam ring 5 becomes zero, and thus the
eccentricity of geometric center "P" of inner peripheral surface 50
of cam ring 5 with respect to the axis "O" of drive shaft 3 (or the
rotation center of rotor 4), that is, the distance |OP| between the
axis "O" and the geometric center "P", becomes maximum. The initial
position (the spring-loaded position) of cam ring 5, held in the
initial setting state, is hereinafter referred to as "initial
setting position". The pump unit is constructed by pump housing 1,
drive shaft 3, rotor 4, cam ring 5, inlet port 16b, discharge port
17b, and vanes 6a-6g. One pump working chamber is defined between
two adjacent vanes 6. That is, seven pump chambers r1-r7 are
defined as seven internal spaces partitioned in a fluid-tight
fashion and surrounded by vanes 6a-6g, two opposed sidewalls (i.e.,
pump-cover bottom face 20a and pump-housing bottom face 10a), rotor
outer peripheral surface 42a, cam-ring inner peripheral surface
50.
[0104] The geometric center "P" of cam-ring inner peripheral
surface 50 is offset from the rotation center "O" of rotor 4 in the
positive y-axis direction in the initial setting state of cam ring
5, shown in FIG. 5. Hence, pump chambers r1, r2, r3, and r4,
defined on the side of the positive x-axis direction with respect
to the center "O", are configured such that respective volumes of
pump chambers r1, r2, r3, and r4 gradually increase, in that order,
i.e., from the side of the negative y-axis direction to the side of
the positive y-axis direction. In a similar manner, pump chambers
r4, r5, r6, and r7, defined on the side of the negative x-axis
direction with respect to the center "O", are configured such that
respective volumes of pump chambers r4, r5, r6, and r7 gradually
decrease, in that order, i.e., from the side of the positive y-axis
direction to the side of the negative y-axis direction. As viewed
from the positive z-axis direction in FIG. 5, drive shaft 3 rotates
rotor 4 in the counterclockwise direction. In the initial setting
state, due to counterclockwise rotation of rotor 4, the volume of
each of pump chambers r1, r2, r3, and r4 tends to increase, while
the volume of each of pump chambers r4, r5, r6, and r7 tends to
decrease.
[0105] As viewed in the z-axis direction, the angle between the
opposed faces of two adjacent vanes (6a-6b, 6b-6c, 6c-6d, 6d-6e,
6e-6f, 6f-6g) is dimensioned to be slightly less than the angle
.angle.AOC or the angle .angle.BOD (see FIG. 4). In other words,
the circumferential distance between the opposed faces of two
adjacent vanes is dimensioned to be slightly less than the
circumferential distance between circumferential end "A" of
circular-arc shaped inlet port 16b and circumferential end "C" of
circular-arc shaped groove 17c of discharge port 17b (or the
circumferential distance between circumferential end "D" of
circular-arc shaped groove 17c of discharge port 17b and
circumferential end "B" of circular-arc shaped inlet port 16b). For
instance, as can be appreciated from comparison of pump chamber r4
shown in FIG. 5 with the area defined between circumferential ends
"A" and "C" shown in FIG. 4, each of pump chambers r1-r7 cannot
overlap simultaneously with both of two opposed circumferential
ends "A" and "C" (or both of two opposed circumferential ends "D"
and "B"). Hence, there is no possibility that each of pump chambers
r1-r7 is communicated simultaneously with both of inlet port 16b
and discharge port 17b. Because of a very small offset between the
axis "O" and the geometric center "P", each of pump chambers r1-r7
cannot be communicated simultaneously with both of inlet port 16b
and discharge port 17b, regardless of the oscillated amount of cam
ring 5, in other words, regardless of the maximum eccentricity (the
initial setting position) or minimum eccentricity of cam ring
5.
[0106] Pump chambers r1, r2, and r3 (further including pump chamber
r4 just before rotor 4 reaches the angular position shown in FIG.
5), defined on the side of the positive x-axis direction, are
configured or defined to overlap with inlet port 16b as viewed from
the z-axis direction, so as to communicate inlet port 16b. In other
words, inlet port 16b is configured to formed in bottom face 10a of
pump housing 1 and configured to open into pump chambers r1-r3 (or
r1-r4) in such a manner as to extend over pump chambers r1-r3 (or
r1-r4) within the area where the volumes of these pump chambers
tend to gradually increase depending on rotary motion
(counterclockwise rotation) of rotor 4. On the other hand, pump
chambers r5, r6, and r7 (further including pump chamber r4 just
after rotor 4 has reached the angular position shown in FIG. 5),
defined on the side of the negative x-axis direction, are
configured or defined to overlap with discharge port 17b as viewed
from the z-axis direction, so as to communicate discharge port 17b.
In other words, discharge port 17b is formed in pump-housing bottom
face 10a and configured to open into pump chambers r5-r7 (or r4-r7)
in such a manner as to extend over pump chambers r5-r7 (or r4-r7)
within the area where the volumes of these pump chambers tend to
gradually decrease depending on rotary motion (counterclockwise
rotation) of rotor 4.
[0107] For the reasons discussed above, pump chambers r1-r3 (or
r1-r4), defined on the side of the positive x-axis direction with
respect to the rotation center "O" of rotor 4, are operating on the
suction stroke (intake stroke), during counterclockwise rotation of
rotor 4. On the other hand, pump chambers r5-r7 (or r4-r7), defined
on the side of the negative x-axis direction with respect to the
rotation center "O" of rotor 4, are operating on the discharge
stroke, during counterclockwise rotation of rotor 4.
[0108] Working oil, discharged from discharge port 17b, is
introduced into back-pressure chambers 40a-40g of rotor 4, thereby
forcing each of vanes 6a-6g radially outwards. Additionally, during
rotation of rotor 4, vanes 6a-6g themselves are forced radially
outwards due to centrifugal force. Hence, during operation of the
engine, the tip of each of vanes 6a-6g is brought into
abutted-engagement (or sliding-contact) with inner peripheral
surface 50 of cam ring 5. In the engine stopped state where there
is no rotation of pump VP, vane rings 7a-7b support vanes 6a-6g so
as to force them radially outwards. By virtue of the supporting
action of vane rings 7a-7b, even during the early stages of engine
starting, it is possible to rapidly ensure a fluid-tight
performance of each of pump chambers r1-r7, thus enhancing a
responsiveness of pump discharge pressure (a rapid discharge
pressure rise). Additionally, by virtue of the supporting action of
vane rings 7a-7b, it is possible to suppress vanes 6a-6g from being
brought into collision-contact with cam-ring inner peripheral
surface 50 owing to radially-outward movements of vanes 6a-6g out
of respective slits 4a-4g, even when pump VP begins to rotate.
[0109] A proper clearance space CL is defined between an outer
peripheral surface 50a of cam ring 5 and inner peripheral surface
13a of pump-housing peripheral wall 13, so as to permit oscillating
motion of cam ring 5. Cam-ring outer peripheral surface 50a is kept
out of contact with pump-housing inner peripheral surface 13a,
except for arm portion 5d. For the reasons discussed later, notice
that, in the variable displacement vane pump VP of the shown
embodiment, any seal member is not installed in clearance space CL.
Thus, clearance space CL is not partitioned with any seal member.
Clearance space CL communicates with the oil pan via inlet hole
16a. Hence, a pressure in clearance space CL around the entire
circumference of cam-ring outer peripheral surface 50a (i.e., a
pressure applied to the outer periphery of cam ring 5) becomes
atmospheric pressure. For this reason, during operation of pump VP,
the pressure in clearance space CL becomes less than a discharge
pressure (denoted by "Pd") of working oil discharged from discharge
port 17b.
[0110] As set forth above, there is a less pressure difference
between the pressure (i.e., atmospheric pressure) outside of pump
housing 1 and the pressure in clearance space CL, and thus it is
unnecessary to interleave a seal member (a gasket usually used to
enhance a fluid-tight performance of pump VP) between pump-housing
flanged portion 14 and pump-cover flanged portion 24. Additionally,
the same pressure (atmospheric pressure) is applied around the
entire circumference of cam-ring outer peripheral surface 50a. That
is, the external pressure, acting on the outer periphery of cam
ring 5 in the direction perpendicular to the axis "O" of drive
shaft 3, becomes approximately uniform. Because of the
approximately uniform outside pressure acting on the cam-ring outer
periphery, there is no occurrence of oscillating motion of cam ring
5, resulting from the external pressure. Therefore, a force, which
creates an oscillating motion (an angular displacement or a pivotal
motion) of cam ring 5 about the pivot (pivot pin 9), can be stably
applied to cam ring 5 via cam-ring inner peripheral surface 50.
[0111] (Construction of Cam Ring)
[0112] The axial width of cam ring 5 (i.e., the length of cam ring
5 in the z-axis direction) is the same around the entire
circumference. The radial width of cam-ring cylindrical portion 5a
partly differs. That is, the radial wall thickness of the lower
half of cylindrical portion 5a (on the side of the negative y-axis
direction with respect to geometric center "P" of cam-ring inner
peripheral surface 50) is dimensioned to be greater than that of
the upper half of cylindrical portion 5a (on the side of the
positive y-axis direction with respect to geometric center "P").
Concretely, a radial width (a radial wall thickness) L2 of the
lower half of cylindrical portion 5a (in particular, a part of the
lower half of cylindrical portion 5a overlapping with inlet port
16b and discharge port 17b on the side of the negative y-axis
direction with respect to geometric center "P" of cam-ring inner
peripheral surface 50) is dimensioned to be greater than a radial
width (a radial wall thickness) L1 of the upper half of cylindrical
portion 5a (on the side of the positive y-axis direction with
respect to geometric center "P").
[0113] That is, as viewed from the z-axis direction, a part of
cylindrical portion 5a, overlapping with inlet port 16b and
discharge port 17b, on the side of the negative y-axis direction
with respect to geometric center "P", is formed to be comparatively
thick-walled, but the radial width (the radial wall thickness) of
the intermediate part of cam-ring cylindrical portion 5a of the
lower half, interconnecting the inlet-port side thick-walled part
and the discharge-port side thick-walled part, is dimensioned to be
identical to the radial width L1 (<L2) of the upper half of
cylindrical portion 5a. That is, the intermediate part is formed as
a recessed portion sandwiched between these thick-walled parts. The
angle between the line segment between and including center "O" and
the clockwise end of the recessed intermediate part and the line
segment between and including center "O" and the counterclockwise
end of the recessed intermediate part is dimensioned to be less
than the previously-described angle .angle.BOD. In other words, as
seen in FIG. 5, the thick-walled part of the side of inlet port 16b
is configured to somewhat extend clockwise by a predetermined
distance from the circumferential end "B" of inlet port 16b. In a
similar manner, the thick-walled part of the side of discharge port
17b is configured to somewhat extend counterclockwise by a
predetermined distance from the circumferential end "D" of
discharge port 17b.
[0114] In the initial setting position of cam ring 5, shown in FIG.
5, a proper radial width (a proper radial distance) L3 between
outer peripheral surface 50a of cam ring 5 and the
circumferentially-extending outside edged portion of each of inlet
port 16b and discharge port 17b can be ensured in a manner so as to
be substantially equal to the radial width L1, around the entire
circumference of cam ring 5. For instance, regarding the left-hand
side (on the side of discharge port 17b) thick-walled portion and
the right-hand side (on the side of inlet port 16b) thick-walled
portion, a radial width (a radial distance) L3 between cam-ring
outer peripheral surface 50a and the circumferentially-extending
outside edged portion of each of inlet port 16b and discharge port
17b is dimensioned to be substantially equal to the radial width
L1, i.e., L3.apprxeq.L1. Also, regarding the previously-discussed
recessed intermediate part of the cam-ring lower half, a distance
from cam-ring outer peripheral surface 50a to the
circumferentially-extending outside edged portion (or the
circumferential end "B", "D") of each of inlet port 16b and
discharge port 17b is dimensioned to be greater than or equal to
the radial width L1.
[0115] Pivot portion 5c is arranged on the outer periphery of
cylindrical portion 5a (on the side of the negative x-axis
direction of cylindrical portion 5a) in such a manner as to be
slightly offset from the x-axis in the positive y-axis direction.
Pivot portion 5c has a small annular shape in lateral cross section
and has pivot bore 52 formed at its center (identical to geometric
center "Q" of pin insertion hole 12). The outer periphery of pivot
portion 5c, facing in the positive y-axis direction, is contoured
or formed as a small circular arc 51b whose center is the geometric
center "Q" of pivot bore 52. As viewed from the z-axis direction,
the small circular-arc shaped outer periphery of pivot portion 5c
is contoured to be substantially conformable to the shape of the
outer periphery of support portion 12a of pump housing 1, facing in
the positive y-axis direction.
[0116] Sector portion 5b is arranged on the side of the positive
y-axis direction of pivot portion 5c. Communication hole 51, formed
to penetrate sector portion 5b, is contoured or configured to be
substantially conformable to the shape of sector groove 17d of
discharge port 17b. As viewed from the z-axis direction, the cross
section of communication hole 51 is set to be greater than or equal
to the cross section of discharge hole 17a. Communication hole 51
is configured to be surrounded in all directions by a large
circular arc 51a whose center is the geometric center "Q" of pivot
bore 52, a small circular arc 51b whose center is the geometric
center "Q", a side 51c substantially parallel to the y-axis, and a
circular arc 51d constructing a part of the outer peripheral
surface of cam-ring cylindrical portion 5a. Large circular arc 51a
is located on the side of the positive y-axis direction of
communication hole 51, whereas small circular arc 51b is located on
the side of the negative y-axis direction of communication hole 51
so as to be opposed to large circular arc 51a. The side 51c is
located on the side of the negative x-axis direction of
communication hole 51, whereas circular arc 51d is located on the
side of the positive x-axis direction of communication hole 51.
[0117] As viewed from the z-axis direction, small circular arc 51b
is configured to be conformable to small circular arc 17f of sector
groove 17d of discharge port 17b. Large circular arc 51a is
configured to be conformable to large circular arc 17e of
discharge-port sector groove 17d. In the initial setting position
of cam ring 5, the side 51c is configured to be conformable to the
side 17g of sector groove 17d.
[0118] When cam ring 5 oscillates or displaces clockwise from the
initial setting position of FIG. 5, sector portion 5b also rotates
clockwise about pivot bore 52. At this time, large circular arc 51a
of communication hole 51 relatively moves on large circular arc 17e
of sector groove 17d, such that the locus of motion of large
circular arc 51a traces large circular arc 17e as viewed from the
z-axis direction. In a similar manner, small circular arc 51b of
communication hole 51 relatively moves on small circular arc 17f of
sector groove 17d, such that the locus of motion of small circular
arc 51b traces small circular arc 17f as viewed from the z-axis
direction. The side 51c of communication hole 51 moves apart from
the side 17g of sector groove 17d and rotates clockwise about pivot
bore 52. As viewed from the z-axis direction, the side 51c is
configured to be kept in a slightly-counterclockwise-spaced,
contact-free relationship with the circumference of discharge hole
17a or kept at the very limit of contact with the circumference of
discharge hole 17a, even when the oscillated amount of cam ring 5
becomes maximum, that is, even with cam ring 5 just displaced to
the minimum-eccentricity position shown in FIG. 11. In other words,
the side 51c of communication hole 51 and discharge hole 17a cannot
be overlapped with each other, regardless of the oscillated amount
of cam ring 5, that is, regardless of the maximum eccentricity (see
the initial setting position (the spring-offset position) of cam
ring 5 shown in FIG. 5) or the minimum eccentricity (see the FIG.
11).
[0119] In an assembled state of cam ring 5 installed in pump
housing 1, communication hole 51 of cam ring 5 serves to
intercommunicate discharge port 17b (sector groove 17d) of pump
housing 1 and discharge port 23 (sector groove 23d) of pump cover
2. During operation of pump VP, most of high-pressure working fluid
(high-pressure working oil), which is supplied from pump chambers
r5-r7 (or pump chambers r4-r7) to discharge port 23 of pump cover
2, is discharged from discharge hole 17a via communication hole
51.
[0120] Now, assume that only a pump-housing discharge port, such as
discharge port 17b, is provided. In such a case, the fluid pressure
in discharge port 17b acts on cam ring 5 in the positive z-axis
direction. In such a case, the fluid pressure forces cam ring 5
toward pump cover 2, and thus the frictional force created between
cam ring 5 and pump cover 2 becomes great. This means an
undesirably large force for oscillating motion of cam ring 5, i.e.,
an undesirably large energy loss. To avoid this, a pump-cover
discharge port, such as discharge port 23, is further provided in
order for the fluid pressure in discharge port 23 to act on cam
ring 5 in the negative z-axis direction, thereby enabling cam ring
5 to be forced apart from pump cover 2.
[0121] However, in the first embodiment, discharge hole 17a is laid
out outside of cam-ring cylindrical portion 5a rather than inside
of cam-ring inner peripheral surface 50 (i.e., the side of the
defined pump chambers) and formed in bottom face 10a (sector groove
17d) of pump housing 1. On the other hand, there are no discharge
holes formed in pump cover 2. Therefore, assuming that
communication hole 51 is not formed in cam ring 5, there is an
increased tendency for working oil to stay in discharge port 23 of
pump cover 2, and thus there is a risk of contaminant and debris
accumulated in pump-cover discharge port 23. Generally, the fluid
pressure applied from the side of pump-cover discharge port 23 to
cam ring 5, tends to be slightly greater than the fluid pressure
applied from the side of pump-housing discharge port 17b to cam
ring 5. Due to the applied fluid-pressure difference, cam ring 5
tends to be forced toward pump housing 1, and thus the frictional
force created between cam ring 5 and pump housing 1 tends to become
great. This means an undesirably large force required for
oscillating motion of cam ring 5, i.e., an undesirably large energy
loss.
[0122] In the first embodiment, pump housing 1 and pump cover 2
have respective discharge ports 17b and 23, and cam ring 5 has
communication hole 51 through which discharge port 17b (sector
groove 17d) of pump housing 1 and discharge port 23 (sector groove
23d) of pump cover 2 are communicated with each other. Hence,
working oil in discharge port 23 (sector groove 23d) of pump cover
2 can flow via communication hole 51 into discharge port 17b
(sector groove 17d) of pump housing 1, and then the working fluid
introduced into discharge port 17b can be discharged from discharge
hole 17a.
[0123] Therefore, in the case of the variable displacement vane
pump unit of the first embodiment having the specific
discharge-hole layout that discharge hole 17a is laid out outside
of cam-ring cylindrical portion 5a rather than inside of cam-ring
inner peripheral surface 50 and formed in bottom face 10a (sector
groove 17d) of pump housing 1, it is possible to increase the
number of working-oil passages communicating with discharge hole
17a, as compared to a typical vane pump unit that any cam-ring
communication hole is not formed, thereby increasing a discharge of
working oil from pump VP (i.e., a fluid flow rate per one
revolution of vane-pump rotor 4), in other words, a pump
discharging effect. Additionally, there is a less risk of
contaminant and debris accumulated in pump-cover discharge port 23.
The fluid pressure applied from the side of pump-housing discharge
port 17b to cam ring 5 and the fluid pressure applied from the side
of pump-cover discharge port 23 to cam ring 5 are almost balanced
with each other, and thus it is possible to hold cam ring 5
substantially at an intermediate position between pump housing 1
and pump cover 2 in the z-axis direction. Hence, it is possible to
reduce or minimize the magnitude of frictional force created
between cam ring 5 and each of pump housing 1 and pump cover 2,
thereby effectively reducing a force required for oscillating
motion of cam ring 5.
[0124] In the shown embodiment, the fluid-flow passage area of
communication hole 51 is dimensioned or set to be greater than or
equal to that of discharge hole 17a, and thus it is possible to
reduce the flow resistance to working-oil flow, caused by
communication hole 51. Communication hole 51 cannot serve as a
fluid-flow constriction orifice, and thus it is possible to
increase the amount of working oil discharged through communication
hole 51 as much as possible, thereby ensuring the increased pump
discharging effect. Furthermore, it is possible to bring the fluid
pressure applied from the side of pump-cover discharge port 23 to
cam ring 5 closer to the fluid pressure applied from the side of
pump-housing discharge port 17b to cam ring 5.
[0125] As set forth above, communication hole 51 of cam ring 5 is
contoured or configured to be substantially conformable to both the
shape of sector groove 17d of pump housing 1 and the shape of
sector groove 23d of pump cover 2. Communication hole 51 is formed
into a circular-arc shape (or a sector form) whose center is the
geometric center "Q" of pivot bore 52 (i.e., the fulcrum "Q" of
oscillating motion of cam ring 5). Hence, even when cam ring 5 is
pivoting about the fulcrum "Q", there is a less change in the
overlapping area between pump-housing sector groove 17d (pump-cover
sector groove 23d) and communication hole 51, in other words, there
is a less change in flow passage cross-sectional area of the
working-oil flow passage oriented from pump-cover discharge port 23
(sector groove 23d) toward pump-housing discharge port 17b. In this
manner, even when cam ring 5 is oscillating, pump-housing sector
groove 17d and pump-cover sector groove 23d are permanently
communicated with each other via communication hole 51, without any
rapid change in the fluid-flow passage area of cam-ring
communication hole 51. Accordingly, it is possible to stably
provide the advantageous operation and effects as described
previously.
[0126] Moreover, as set forth above, the side 51c of communication
hole 51 is kept outside of discharge hole 17a or kept at the very
limit of contact with the circumference of discharge hole 17a but
not overlap with discharge hole 17a, during oscillating motion of
cam ring 5. Thus, there is no change in the opening area of
communication hole 51 opening into discharge hole 17a, that is,
there is no change in fluid-flow passage cross-sectional area of
the working-fluid flow passage oriented from pump-cover discharge
port 23 (sector groove 23d) toward pump-housing discharge hole 17a.
Hence, there is a less change in the fluid pressure applied from
the side of pump-cover discharge port 23 to cam ring 5. Hence, it
is possible to stably provide the advantageous operation and
effects as described previously.
[0127] Cam-ring arm portion 5d is formed of a substantially
rectangular support portion 53 overhanging from cam-ring
cylindrical portion 5a in the positive x-axis direction and a
protruding portion 54 having a substantially semicircular cross
section and extending downwards from the underside of support
portion 53, facing in the negative y-axis direction. In the initial
setting state of cam ring 5, shown in FIG. 5, the upside of support
portion 53, facing in the positive y-axis direction, is kept in
wall-contact with seat surface 15b (of arm-portion accommodating
chamber 15a) on which arm portion 5d of cam ring 5 is seated.
[0128] The surface 54a of protruding portion 54 is formed as a
curved surface. As viewed from the z-axis direction, protruding
portion 54 is formed into a semicircle in cross section. In the
initial setting state shown in FIG. 5, the lowermost end of
protruding portion 54, facing in the negative y-axis direction, is
laid out substantially at the same level as shoulder portions
15h-15i of spring chamber 15d. As viewed in the x-axis direction,
the center of the lowermost end (viewing FIG. 5) of protruding
portion 54 is aligned with the neutral axis of spring chamber 15d.
The maximum width of protruding portion 54, measured in the x-axis
direction, is dimensioned to be less than the maximum width of the
opening of spring chamber 15d, measured in the x-axis direction, in
other words, the distance between the opposed shoulder portions
15h-15i.
[0129] (Construction of Biasing Member)
[0130] Biasing member 8 has a double spring structure, in which a
first coil spring 8a is coaxially installed inside of a second coil
spring 8b. FIG. 6 shows the cross section of pump VP whose pump
cover is installed, taken along the line E-E of FIG. 5. First coil
spring 8a is installed inside of second coil spring 8b, such that
the coiled directions (or winding directions) of first and second
coil springs 8a-8b are opposite to each other.
[0131] The coil outside diameter of first coil spring 8a is
dimensioned to be less than the maximum width of the opening of
spring chamber 15d, measured in the x-axis direction, in other
words, the distance between the opposed shoulder portions 15h-15i,
and also dimensioned to be substantially equal to the width of
protruding portion 54, measured in the x-axis direction. The coil
outside diameter of second coil spring 8b is dimensioned to be
substantially equal to the width of spring chamber 15d, measured in
the x-axis direction (see FIG. 5), and also dimensioned to be less
than the length of protruding portion 54 (or the length of each of
shoulder portions 15h-15i of spring chamber 15d), measured in the
z-axis direction. The lower coil end of each of first and second
coil springs 8a-8b, facing in the negative y-axis direction is
seated on the bottom face 15e of spring chamber 15d. The
diametrically-opposing coil-end portions (opposed to each other in
the z-axis direction) of the upper coil end of first coil spring
8a, facing in the positive y-axis direction, are kept in
abutted-engagement with the lowermost end of protruding portion 54,
facing in the negative y-axis direction, without engagement between
first coil spring 8a and shoulder portions 15h-15i. First coil
spring 8a is installed in spring chamber 15d and disposed between
pump housing 1 (i.e., spring-chamber bottom face 15e) and cam-ring
arm portion 5d (i.e., protruding portion 54) under a preloaded
condition where first coil spring 8a is preloaded by an initial set
load W1.
[0132] On the other hand, the upper coil end of second coil spring
8b, facing in the positive y-axis direction, is engaged with
shoulder portions 15h-15i. The diametrically-opposing coil-end
portions (opposed to each other in the x-axis direction) of the
upper coil end of second coil spring 8b, facing in the positive
y-axis direction, are kept in abutted-engagement with the
respective undersides of shoulder portions 15h-15i, facing in the
negative y-axis direction. Second coil spring 8b is installed in
spring chamber 15d and disposed between pump housing 1 (i.e.,
spring-chamber bottom face 15e) and the shoulder pair 15h-15i under
a preloaded condition where second coil spring 8b is preloaded by
an initial set load W3.
[0133] (Operation Carried Out By Layout of Fulcrum of Oscillating
Motion of Cam Ring)
[0134] Hereinafter described is the operation, carried out by a
specific layout of the fulcrum of oscillating motion of cam ring 5.
As discussed above, the fulcrum of cam ring 5 is the geometric
center "Q" of pin insertion hole 12 (in other words, the geometric
center "Q" of pivot bore 52 or the geometric center "Q" of pivot
pin 9). The fulcrum "Q" of cam ring 5 is laid out to be offset in
the biasing direction of biasing member 8 (i.e., in the positive
y-axis direction) within an opening range of discharge port 17b.
That is, the fulcrum "Q" of cam ring 5 is laid out to be offset in
the biasing direction of biasing member 8 (i.e., in the positive
y-axis direction) with respect to a midpoint (a center position) of
the opening range of discharge port 17b. In other words, pivot bore
52 formed in cam ring 5, is laid out or configured, such that an
area of cam-ring inner peripheral surface 50, on which the fluid
pressure in discharge port 17b acts during operation of pump VP and
which is segmented as a second pressure-receiving area S2
(described later in detail in reference to the chart of FIG. 9)
extending in the positive y-axis direction (in the biasing
direction of biasing member 8) with respect to the fulcrum "Q" of
oscillating motion of cam ring 5 is permanently (always) smaller
than an area of cam-ring inner peripheral surface 50, on which the
fluid pressure in discharge port 17b acts during operation of pump
VP and which is segmented as a first pressure-receiving area S1
(described later in detail in reference to the chart of FIG. 9)
extending in the negative y-axis direction (in the direction
opposite to the biasing direction of biasing member 8) with respect
to the fulcrum "Q". The details of the specific layout of the
fulcrum "Q" of oscillating motion of cam ring 5 are described
hereunder.
[0135] FIGS. 7-8 are the partial views (component parts shown in
FIG. 5 are partly omitted), for explaining two different angular
positions of oscillating or rotating motion of cam ring 5 relative
to pump housing 1. FIG. 7 shows the initial setting state (the
maximum-eccentricity state) where the degree of oscillating or
rotating motion of cam ring 5 relative to pump housing 1 is
minimum, whereas FIG. 8 is the minimum-eccentricity state where the
degree of oscillating or rotating motion of cam ring 5 relative to
pump housing 1 is maximum.
[0136] As viewed from the z-axis direction, the point that the
outside edged portion of circular-arc shaped groove 17c of
discharge port 17b intersects (overlaps) with cam-ring inner
peripheral surface 50 at the circumferential end "C" of
discharge-port circular-arc shaped groove 17c, facing in the
positive y-axis direction, is defined as a point "C'". The point
that the outside edged portion of discharge-port circular-arc
shaped groove 17c intersects (overlaps) with cam-ring inner
peripheral surface 50 at the circumferential end "D" of
discharge-port circular-arc shaped groove 17c, facing in the
negative y-axis direction, is defined as a point "D'".
Additionally, as viewed from the z-axis direction, the point that
the straight line segment PQ, which links the fulcrum "Q" of
oscillating motion of cam ring 5 and the geometric center "P" of
cam-ring inner peripheral surface 50, intersects with cam-ring
inner peripheral surface 50 on the side of discharge port 17b, is
defined as an intersection point "R". As viewed from the z-axis
direction, the circular-arc segment C'RD' of cam-ring inner
peripheral surface 50 is laid out to overlap with discharge-port
circular-arc shaped groove 17c.
[0137] As can be seen from the two explanatory views of FIGS. 7-8,
the positional relationship between the fulcrum "Q" of oscillating
motion of cam ring 5 and the position of formation of discharge
port 17b is predetermined, such that the arc length of the
circular-arc segment C'R (extending in the positive y-axis
direction with respect to the fulcrum "Q") is less than the arc
length of the circular-arc segment RD' (extending in the negative
y-axis direction with respect to the fulcrum "Q"), regardless of
the presence or absence of oscillating motion of cam ring 5. That
is, assuming that the midpoint of the circular-arc segment C'RD' is
defined as a point "S", the intersection point "R" is laid out to
be offset from the midpoint "S" in the positive y-axis direction
(that is, in the direction r7.fwdarw.r6.fwdarw.r5 in which
respective volumes of pump chambers r7, r6, and r5 gradually
increase, in that order), over the entire range of oscillating
motion of cam ring 5. In other words, the intersection point "R" is
laid out to be offset from the midpoint "S" in the
maximum-eccentricity direction, that is, in the direction that the
geometric center "P" of cam-ring inner peripheral surface 50
displaces from the axis "O" of drive shaft 3 (i.e., the rotation
center "O" of vane rotor 4).
[0138] The fluid pressure (discharge pressure Pd of a high pressure
level) in discharge port 17b acts on the circular-arc segment C'RD'
of cam-ring inner peripheral surface 50 (the inner peripheral
surface on which vanes 6a-6g slide), which segment C'RD' overlaps
with discharge port 17b. As discussed above, in the variable
displacement vane pump construction of the first embodiment, there
is a difference between the second segmented pressure-receiving
area (i.e., the circular-arc segment C'R of cam-ring inner
peripheral surface 50) extending in the positive y-axis direction
with respect to the intersection point "R" (serving as a boundary
point or a reference point) and the first segmented
pressure-receiving area (i.e., the circular-arc segment RD' of
cam-ring inner peripheral surface 50) extending in the negative
y-axis direction with respect to the intersection point "R". Due to
the aforementioned difference between the first and second
segmented pressure-receiving areas, created by the specific layout
of the fulcrum "Q" of oscillating motion of cam ring 5, it is
possible to produce a moment by which cam ring 5 can be rotated or
oscillated about the fulcrum "Q" against the spring bias of biasing
member 8.
[0139] In the explanatory views of FIGS. 7-8, a vector Fa denotes a
force produced by discharge pressure Pd received by cam-ring inner
peripheral surface 50, ranging from the intersection point "R" to
the circumferential end "D" of discharge-port circular-arc shaped
groove 17c. On the other hand, a vector Fb denotes a force produced
by discharge pressure Pd received by cam-ring inner peripheral
surface 50, ranging from the intersection point "R" to the
circumferential end "C" of discharge-port circular-arc shaped
groove 17c. As a result of force Fa, a clockwise moment Ta of that
force Fa about the fulcrum "Q" of oscillating motion of cam ring 5
is produced. As a result of force Fb, a counterclockwise moment Tb
of that force Fb about the fulcrum "Q" of oscillating motion of cam
ring 5 is produced.
[0140] In both the maximum-eccentricity state (the initial setting
state) shown in FIG. 7 and the minimum-eccentricity state (the
zero-eccentricity state) shown in FIG. 8, the arc length of
circular-arc segment RD' of cam-ring inner peripheral surface 50 is
longer than that of circular-arc segment C'R. Hence, the force Fa
becomes greater than the force Fb, and thus the magnitude of moment
Ta becomes greater than that of moment "Tb", i.e., Ta>Tb. In
other words, due to the difference between two forces Fa and Fb of
different magnitude and opposite sense (i.e., the difference
between two moments Ta and Tb of different magnitude and opposite
rotation direction), the moment of force about the fulcrum "Q" of
oscillating motion of cam ring 5 and having the magnitude (Ta-Tb)
and clockwise rotation direction is produced totally, so as to
reduce the eccentricity |OP| of geometric center "P" of cam-ring
inner peripheral surface 50 with respect to the axis "O" of drive
shaft 3 (or the rotation center of vane rotor 4).
[0141] Such a moment Ta-Tb, resulting from discharge pressure Pd
acting on cam-ring inner peripheral surface 50, is created by the
offset layout of the intersection point "R" with respect to the
midpoint "S". That is, the moment Ta-Tb is created by the specific
cam-ring oscillating-motion fulcrum layout that the intersection
point "R" is laid out to be offset from the midpoint "S" in the
direction (i.e., in the positive y-axis direction) that the
eccentricity |OP| of geometric center "P" with respect to the axis
"O" increases. In other words, the moment Ta-Tb is created by the
specific cam-ring oscillating-motion fulcrum layout that the
fulcrum "Q" of oscillating motion of cam ring 5 is offset from the
midpoint "S" in the biasing direction (i.e., in the positive y-axis
direction) of biasing member 8.
[0142] The position of midpoint "S" of the circular-arc segment
C'RD' is determined by the layout of cam ring 5 relative to
discharge port 17b. In the shown embodiment, the radial width of
discharge port 17b (circular-arc shaped groove 17c), the inside
diameter of cam-ring inner peripheral surface 50, and the layout of
cam ring 5 relative to discharge port 17b (circular-arc shaped
groove 17c) are set, dimensioned, and configured, such that, as
viewed from the z-axis direction, there is a less change of the
overlapping area that cam-ring inner peripheral surface 50 overlaps
with both discharge port 17b (circular-arc shaped groove 17c) and
inlet port 16b, during oscillating motion of cam ring 5.
[0143] In other words, cam-ring inner peripheral surface 50 is laid
out or configured so as to be able to oscillate within the
designated area that the overlapping of cam-ring inner peripheral
surface 50 with both the circumferential ends "C" and "D" of
discharge-port circular-arc shaped groove 17c is permitted during
oscillating motion of cam ring 5. That is, during oscillating
motion of cam ring 5, the point "C'" merely moves on the
circumferential end "C", while the point "D'" merely moves on the
circumferential end "D". During oscillating motion of cam ring 5,
there is a less change in the position of each of points "C'" and
"D'", and thus there is a less change in the position of the
midpoint "S" of circular-arc segment C'RD'. In other words, a
change in the position of midpoint "S", occurring during
oscillating motion of cam ring 5, is a negligibly small change by
which the length of circular-arc segment SR is unaffected.
[0144] Therefore, the position of the midpoint "S" on cam-ring
inner peripheral surface 50 can be approximated to the position of
the circumferential midpoint of discharge port 17b (circular-arc
shaped groove 17c). That is to say, it will be understood that the
fulcrum "Q" of oscillating motion of cam ring 5 is laid out to be
offset in the biasing direction of biasing member 8 with respect to
the midpoint of discharge port 17b (circular-arc shaped groove
17c), thereby creating the previously-discussed moment Ta-Tb.
[0145] On the assumption that vanes 6a-6g are positioned at their
positions as shown in FIGS. 7-8, the aforementioned moment Ta-Tb,
created by the fluid pressure (discharge pressure Pd) received by a
specific portion (circular-arc segment C'RD') of cam-ring inner
peripheral surface 50 that overlaps with the opening range of
discharge port 17b (circular-arc shaped groove 17c), has been
explained previously. This is because the actual difference between
the first and second segmented pressure-receiving areas can be
sufficiently approximated to the difference between the arc lengths
of circular-arc segments C'R and RD' of cam-ring inner peripheral
surface 50. However, actually, regarding pump chamber r7, discharge
pressure Pd in discharge port 17b is also applied to a slight area
of pump chamber r7, ranging from a portion of cam-ring inner
peripheral surface 50 (with which the tip of vane 6e abuts) to the
point "D'". Furthermore, assuming that vane rotor 4 slightly
rotates counterclockwise from the angular position of vane rotor 4
as shown in FIGS. 7-8, and pump chamber r4 is communicated with
discharge port 17b, discharge pressure Pd in discharge port 17b is
also applied to a slight area of pump chamber r4, ranging from a
portion of cam-ring inner peripheral surface 50 (with which the tip
of vane 6b abuts) to the point "C'".
[0146] Referring to FIG. 9, there is shown a time chart
illustrating both a variation in first pressure-receiving area S1
extending in the negative y-axis direction with respect to the
intersection point "R" and a variation in second pressure-receiving
area S2 extending in the positive y-axis direction with respect to
the intersection point "R", during operation of pump VP. For
convenience's sake, assume that each of first and second
pressure-receiving areas S1-S2 is represented simply by the number
of pump chambers, which can be communicated with discharge port 17b
and through which discharge pressure Pd can be applied to cam-ring
inner peripheral surface 50. The force Fa is represented by the
equality Fa=(S1.times.Pd), whereas the force Fb is represented by
the equality Fb=(S2.times.Pd). Thus, on the assumption that
discharge pressure Pd is uniform, the ratio Fa/Fb of two forces Fa
and Fb (in other words, the ratio Ta/Tb of two moments Ta and Tb),
is equal to the ratio S1/S2 between first and second
pressure-receiving areas S1-S2.
[0147] At the time t0, vanes 6a-6g are positioned at the positions
shown in FIGS. 7-8. Pump chambers, receiving discharge pressure Pd
on the side of the negative y-axis direction with respect to the
intersection point "R", are two pump chambers r6, r7, and thus
S1=2. As vane rotor 4 rotates counterclockwise from the time t0,
regarding pump chamber r7, the distance between a portion of
cam-ring inner peripheral surface 50 (with which the tip of vane 6e
abuts) and the point "D'" tends to gradually increase. Immediately
before the time t1 when vane 6e reaches the circumferential end "B"
of inlet port 16b, pump chambers, receiving discharge pressure Pd
on the side of the negative y-axis direction with respect to the
intersection point "R", are half of pump chamber r5, and two pump
chambers r6, r7, and thus first pressure-receiving area S1 becomes
the maximum (=2.5), that is, S1=2.5. As soon as vane 6e reaches the
circumferential end "B" of inlet port 16b at the time t1, fluid
communication between pump chamber r7 and discharge port 17b
becomes blocked, and as a result the number of pumps, receiving
discharge pressure Pd, is decremented by "1", that is, S1=1.5. With
vane rotor 4 rotating counterclockwise, just after the time t1,
regarding pump chamber r6, the distance between a portion of
cam-ring inner peripheral surface 50 (with which the tip of vane 6f
abuts) and the point "D'" tends to gradually increase, and thus
first pressure-receiving area S1 also tends to increase. At the
time t2 when vane 6f reaches the midpoint between the
circumferential end "D" of discharge port 17b and the
circumferential end "B" of inlet port 16b, the number of pump
chambers, receiving discharge pressure Pd on the side of the
negative y-axis direction with respect to the intersection point
"R", becomes "2" in a similar manner to the time t0, that is, S1=2.
After the time t2, the number of pump chambers, receiving discharge
pressure Pd on the side of the negative y-axis direction with
respect to the intersection point "R", can be repeatedly varied at
the same cycle as the time interval between t0 and t2, in other
words, at a given cycle T (=t2-t0). The given cycle T is determined
depending on the distance between two adjacent vanes (6a-6b, 6b-6c,
6c-6d, 6d-6e, 6e-6f, 6f-6g), and the revolution speed of pump VP.
Actually, the given period T corresponds to a time duration
required in order for vane rotor 4 to counterclockwise rotate by an
angle corresponding to one pump chamber.
[0148] On the other hand, pump chambers, receiving discharge
pressure Pd on the side of the positive y-axis direction with
respect to the intersection point "R", are only one pump chamber r5
at the time t0, and thus S2=1. When vane rotor 4 slightly rotates
counterclockwise from the angular position shown in FIGS. 7-8,
fluid communication between pump chamber r4 and discharge port 17b
is established. Thus, with a slight elapsed time from the time t0,
pump chambers, receiving discharge pressure Pd on the side of the
positive y-axis direction with respect to the intersection point
"R", become two pump chambers r4, r5, and thus S2=2.
[0149] Thereafter, as vane rotor 4 rotates counterclockwise,
regarding pump chamber r4, the distance between a portion of
cam-ring inner peripheral surface 50 (with which the tip of vane 6b
abuts) and the point "C'" tends to gradually decrease, and thus
second pressure-receiving area S2 also tends to decrease. At the
time t1 when vane 6b reaches the midpoint between the
circumferential end "A" of inlet port 16b and the circumferential
end "C" of discharge port 17b, the previously-noted distance
between the abutted portion of cam-ring inner peripheral surface 50
with the tip of vane 6b and the point "C'" becomes half of that
obtained at the time t0. At this time (i.e., at the time t1), pump
chambers, receiving discharge pressure Pd on the side of the
positive y-axis direction with respect to the intersection point
"R", are half of pump chamber r5, and pump chamber r4, and thus
second pressure-receiving area S2 totally becomes "1.5", that is,
S2=1.5. After the time t1, as vane rotor 4 rotates
counterclockwise, regarding pump chamber r4, the distance between
the portion of cam-ring inner peripheral surface 50 (with which the
tip of vane 6b abuts) and the point "C'" tends to further decrease,
and then becomes "O" at the time t2 when vane 6b reaches the
circumferential end "C" of discharge port 17b. At this time (i.e.,
at the time t2), the number of pump chambers, receiving discharge
pressure Pd on the side of the positive y-axis direction with
respect to the intersection point "R", becomes "1" in a similar
manner to the time t0, that is, S2=1. After the time t2, the number
of pump chambers, receiving discharge pressure Pd on the side of
the positive y-axis direction with respect to the intersection
point "R", can be repeatedly varied at the same cycle as the time
interval between t0 and t2, in other words, at a given cycle T
(=t2-t0).
[0150] For the reasons discussed above, as can be appreciated from
the time chart of FIG. 9, the relationship between the magnitudes
of first and second segmented pressure-receiving areas S1-S2,
defined by the inequality (S1-S2)>0, can be achieved always (at
each point of time) during operation of pump VP. As seen from the
right-hand diagonal shading area of FIG. 9, the integral
{.intg.(S1-S2)dt} of the difference (S1-S2) of first and second
segmented pressure-receiving areas S1 and S2 for the given cycle T
can be kept at a value greater than "O", that is,
{.intg.(S1-S2)dt}>0, during operation of pump VP. Actually,
there are slight fluctuations in the difference (S1-S2) of first
and second segmented pressure-receiving areas S1 and S2 at
respective points of time during operation of pump VP. However,
such fluctuations are equalized for the given period T. On the
average, the fluctuations are negligible. Thus, the relationship
between the magnitudes of first and second segmented
pressure-receiving areas S1 and S2 can be regarded as to be given
by the inequality S1>S2 for the given period T during operation
of pump VP. For the reasons discussed above, on the average, the
relationship between the magnitude of the force Fa resulting from
the fluid pressure acting on cam-ring inner peripheral surface 50
corresponding to first segmented pressure-receiving area S1
extending in the negative y-axis direction with respect to the
intersection point "R" and the magnitude of the force Fb resulting
from the fluid pressure acting on cam-ring inner peripheral surface
50 corresponding to second segmented pressure-receiving area S2
extending in the positive y-axis direction with respect to the
intersection point "R" can be regarded as to be given by the
inequality Fa>Fb, during operation of pump VP. The positive
moment (the clockwise moment) Ta-Tb(>0) about the fulcrum "Q",
resulting from these forces Fa and Fb, can be continuously produced
during operation of pump VP. That is, by virtue of the fluid
pressure acting on cam-ring inner peripheral surface 50 and the
offset layout of the fulcrum "Q" of oscillating motion of cam ring
5, it is possible to always produce the torque (i.e., the clockwise
moment Ta-Tb), which causes an oscillating or rotating motion of
cam ring 5 in the direction that the eccentricity |OP| of geometric
center "P" of cam-ring inner peripheral surface 50 with respect to
the axis "O" of drive shaft 3 (or the rotation center of rotor 4)
reduces, during operation of pump VP.
[0151] It will be appreciated that a different type of variable
displacement vane pump enables application of a torque (i.e., a
clockwise moment Ta-Tb), which causes an oscillating or rotating
motion of cam ring 5 in the direction that the eccentricity |OP| of
geometric center "P" of cam-ring inner peripheral surface 50 with
respect to the axis "O" reduces, during operation of pump VP, if,
on the average, the integral {.intg.(S1-S2)dt} of the difference
(S1-S2) of first and second pressure-receiving areas S1 and S2 for
a given cycle T can be kept at a value greater than "O", that is,
{.intg.(S1-S2)dt}>0, even when pump VP has a different design
that a time-varied characteristic of first pressure-receiving area
S1 becomes temporarily below that of second pressure-receiving area
S2 within a specific time zone.
[0152] In other words, as far as first and second
pressure-receiving areas S1 and S2 satisfy the relationship defined
by the inequality {.intg.(S1-S2)dt}>0, even when pump VP has a
different cam-ring fulcrum design/configuration that the fulcrum
"Q" of oscillating motion of cam ring 5 is laid out at the
intermediate position of discharge port 17b (circular-arc shaped
groove 17c) and thus there is no offset between the intersection
point "R" and the midpoint "S" at a certain angular position of cam
ring 5 oscillating, it is possible to apply the torque (i.e., the
clockwise moment Ta-Tb), which causes an oscillating or rotating
motion of cam ring 5 in the direction that the eccentricity |OP| of
geometric center "P" of cam-ring inner peripheral surface 50 with
respect to the axis "O" reduces, during operation of pump VP.
[0153] (Operation of Biasing Member)
[0154] As set out above, biasing member 8 is installed in pump
housing 1 at the position substantially symmetrical to the fulcrum
"Q" of oscillating motion of cam ring 5 with respect to the axis
"O" of drive shaft 3. Biasing member 8 forces cam ring 5 in the
direction (in the counterclockwise direction, viewing FIG. 5) that
the volume difference between the volume of the largest pump
chamber (pump chamber r4 in FIG. 5) of pump chambers r1-r7 and the
volume of the smallest pump chamber (pump chambers r1, r7 in FIG.
5) increases, in other words, in the direction that the rate of
change of the volume of each of pump chambers r1-r7 increases.
[0155] In FIGS. 7-8, a vector Fs denotes a force acting on cam ring
5 and produced by a biasing force (e.g., a spring force) of biasing
member 8. In the first embodiment, concretely, force Fs corresponds
to a force vector acting on cam-ring arm portion 5d in the positive
y-axis direction. Force Fs produces a moment Ts by which cam ring 5
can be rotated about the fulcrum "Q" of oscillating motion in the
counterclockwise direction that the eccentricity |OP| of geometric
center "P" of cam-ring inner peripheral surface 50 with respect to
the axis "O" of drive shaft 3 (or rotor 4) increases. The direction
of oscillating motion and the oscillated or rotated position of cam
ring 5 are determined depending on the relationship between the
magnitude of moment Ts, produced by the spring bias of biasing
member 8, and the magnitude of moment Ta-Tb, resulting from the
fluid pressure (discharge pressure Pd) acting on cam-ring inner
peripheral surface 50. The operation of biasing member 8 is
hereunder described in reference to the front elevation views of
FIGS. 10-11 as well as the front elevation view of FIG. 5. As
previously discussed, FIG. 5 shows the initial setting state (the
maximum-eccentricity state) of cam ring 5, where the amount of
oscillating motion of cam ring 5 becomes zero and thus the
eccentricity |OP| of geometric center "P" of cam-ring inner
peripheral surface 50 with respect to the axis "O" of drive shaft 3
becomes maximum. In contrast, FIG. 11 shows the
minimum-eccentricity state of cam ring 5, where the oscillated
amount of cam ring 5 becomes maximum and thus the eccentricity |OP|
of geometric center "P" of cam-ring inner peripheral surface 50
with respect to the axis "O" of drive shaft 3 becomes minimum
(zero), that is, the axis "O" and the geometric center "P" accord
with each other. On the other hand, FIG. 10 shows the
intermediate-eccentricity holding state of cam ring 5, where the
eccentricity |OP| of geometric center "P" of cam-ring inner
peripheral surface 50 with respect to the axis "O" of drive shaft 3
becomes a substantially intermediate value between the maximum and
minimum eccentricities, and a clockwise moment (described later),
resulting from discharge pressure Pd acting on cam-ring inner
peripheral surface 50, is balanced to a counterclockwise moment
(described later), resulting from a summed spring load of first and
second coil springs 8a-8b, both constructing biasing member 8.
[0156] When the revolution speed of pump VP is low, cam ring 5 is
kept in the initial setting state of FIG. 5. That is, the fluid
pressure (discharge pressure Pd) in discharge port 17b is still low
at low revolution speeds of pump VP, the hydraulic pressure (i.e.,
the force difference Fa-Fb of unbalanced forces Fa, Fb), acting on
cam-ring inner peripheral surface 50 in the direction that the
eccentricity |OP| of geometric center "P" with respect to the axis
"O" reduces, is still small. Thus, regarding the moments about the
fulcrum "Q" of oscillating motion of cam ring 5, the
counterclockwise moment Ts about the fulcrum "Q", produced by the
biasing force of biasing member 8, becomes greater than the
clockwise moment Ta-Tb, resulting from the fluid pressure
(discharge pressure Pd) acting on cam-ring inner peripheral surface
50. Accordingly, cam ring 5 is kept in the maximum-eccentricity
state (the initial setting position of FIG. 5).
[0157] As the pump revolution speed increases and thus discharge
pressure Pd from discharge port 17b builds up, the hydraulic
pressure (i.e., the force difference Fa-Fb of unbalanced forces Fa,
Fb), which pressure causes oscillating motion of cam ring 5 about
the fulcrum "Q" against the biasing force (e.g., the spring force)
of biasing member 8 in the direction (i.e., in the clockwise
direction) that reduces the eccentricity |OP| of geometric center
"P" with respect to the axis "O", gradually increases. When
discharge pressure Pd reaches a predetermined pressure value, the
clockwise moment Ta-Tb, resulting from discharge pressure Pd acting
on cam-ring inner peripheral surface 50, becomes identical to a
counterclockwise moment Ts1 about the fulcrum "Q", produced by only
the spring force of first coil spring 8a of two coil springs 8a-8b,
constructing biasing member 8. When discharge pressure Pd exceeds
the predetermined pressure value, the clockwise moment Ta-Tb,
resulting from discharge pressure Pd, becomes greater than the
counterclockwise moment Ts1, produced by only the spring force of
first coil spring 8a. Thus, cam ring 5 begins to oscillate or
rotate clockwise from the maximum-eccentricity position (the
initial setting position of FIG. 5). As a result, cam-ring arm
portion 5d moves in the negative y-axis direction apart from its
seat surface 15b, while compressing first coil spring 8a.
[0158] By clockwise oscillating or rotating motion of cam ring 5,
protruding portion 54 of cam-ring arm portion 5d displaces from the
opening end of spring chamber 15d, facing in the positive y-axis
direction, toward the inside of spring chamber 15d, while
compressing first coil spring 8a. At this time, as seen in FIG. 10,
the lowermost end of protruding portion 54, facing in the negative
y-axis direction, is brought into abutted-engagement with the upper
coil end of second coil spring 8b, which second coil spring is
retained in spring chamber 15d with abutted-engagement of its upper
coil end with the undersides of shoulder portions 15h-15i. At this
time, the eccentricity of cam ring 5 (i.e., the distance |OP|
between the axis "O" and the geometric center "P") becomes a
substantially intermediate value between the maximum eccentricity
(see FIG. 5) and the minimum eccentricity (see FIG. 11). Such an
intermediate-eccentricity state of cam ring 5 is simply referred to
as "holding state". The position of cam ring 5, which is kept in
its holding state, is hereinafter referred to as "holding
position". The rate of change of the volume of each of pump
chambers r1-r7 in the holding state (see FIG. 10) is lower than
that in the initial setting state (see FIG. 5), and higher than
that in the minimum-eccentricity state (see FIG. 11).
[0159] When discharge pressure Pd is within a predetermined
pressure range, the clockwise moment Ta-Tb, resulting from
discharge pressure Pd, becomes greater than the counterclockwise
moment Ts1, resulting from the spring force of first coil spring
8a, and less than a total counterclockwise moment Ts, which total
counterclockwise moment is expressed as the sum (Ts1+Ts2) of the
moment Ts1, resulting from the spring force of first coil spring
8a, and the moment Ts2, resulting from the spring force of second
coil spring 8b, that is, Ts1<(Ta-Tb)<Ts(=Ts1+Ts2). At this
time, the oscillated position of cam ring 5 relative pump housing 1
remains unchanged and thus cam ring 5 can be kept at the holding
position shown in FIG. 10. As soon as discharge pressure Pd exceeds
the above-mentioned predetermined pressure range, the clockwise
moment Ta-Tb, resulting from discharge pressure Pd, becomes greater
than the total counterclockwise moment Ts(=Ts1+Ts2). As a result,
cam ring 5 begins to further oscillate clockwise from the holding
position. Thus, protruding portion 54 of cam-ring arm portion 5d
further displaces in the negative y-axis direction from the opening
end of spring chamber 15d toward the inside of spring chamber 15d,
while compressing second coil spring 8b as well as first coil
spring 8a.
[0160] When discharge pressure Pd reaches a predetermined high
pressure value, first and second coil springs 8a-8b are further
compressed a given stroke (see FIG. 11), and thus the outermost end
of the underside of support portion 53 (the top end of support
portion 53 on the side of a combined direction of the positive
x-axis direction and the negative y-axis direction of cam-ring arm
portion 5d) is brought into abutted-engagement with the upside of
shoulder portion 15i facing in the positive y-axis direction. Under
these conditions, the shoulder portion 15i serves as a stopper for
restricting or preventing a further clockwise rotation of cam ring
5 about the fulcrum "Q". At this time, as seen in FIG. 11, the
geometric center "P" accords with the axis "O", with the result
that the eccentricity of cam ring 5 (i.e., the distance |OP|
between the axis "O" and the geometric center "P") becomes zero. In
the minimum-eccentricity state (the zero-eccentricity state), the
volume difference between the volume of the largest pump chamber
(pump chamber r4 in FIG. 11) of pump chambers r1-r7 and the volume
of the smallest pump chamber (pump chambers r1, r7 in FIG. 11)
becomes minimum (almost zero). In the shown embodiment, in the
minimum-eccentricity state, the eccentricity |OP| of cam ring 5
becomes zero. It will be understood that the minimum-eccentricity
state is not limited to such a zero-eccentricity state, but that,
in the minimum-eccentricity state, a predetermined slight offset
between the geometric center "P" and the axis "O" may exist.
[0161] Referring now to FIG. 12, there is shown the relationship
between a displacement of biasing member 8 (first and second coil
springs 8a-8b) and a load W, in the case of pump VP of the first
embodiment. Concretely, the displacement of biasing member 8 means
a compressed stroke (a compressive deformation) of the double
spring (constructed by first and second coil springs 8a-8b),
corresponding to an oscillated amount (or an oscillated angle) of
cam ring 5 displaced or oscillated clockwise from the initial
setting position. The load W means a spring load corresponding to a
summed value of the spring load produced by first coil spring 8a
and the spring load produced by second coil spring 8b, i.e., the
total spring bias of first and second coil springs 8a-8b. In other
words, spring load W is regarded as to be equivalent to the
previously-noted total counterclockwise moment Ts, expressed as the
sum (Ts1+Ts2) of the moment Ts1, resulting from the spring force of
first coil spring 8a, and the moment Ts2, resulting from the spring
force of second coil spring 8b.
[0162] In the initial setting position (see FIG. 5) of cam ring 5,
the spring load becomes the initial set load W1 of first coil
spring 8a. While cam ring 5 displaces from the initial setting
position (see FIG. 5) toward the holding position (see FIG. 10),
only the first coil spring 8a is compressed and deformed. Hence,
during this clockwise displacement of cam ring 5 between the
initial setting position and the holding position, the spring load
is in proportion to the displacement of first coil spring 8a from
the initial setting position. Thus, the spring load tends to
increase at a gradient corresponding to the spring constant (the
spring stiffness or the spring rate) of first coil spring 8a.
Immediately before cam ring 5 reaches the holding position (see
FIG. 10), the spring load becomes a load W2 (>W1), corresponding
to the displacement of first coil spring 8a at that time.
Therefore, with cam ring 5 displaced at an angular position between
the initial setting position (FIG. 5) and the holding position
(FIG. 10), the moment Ts, resulting from biasing member 8, becomes
identical to the counterclockwise moment Ts1 about the fulcrum "Q",
produced by only the spring load of first coil spring 8a.
[0163] As soon as cam ring 5 further rotates clockwise and reaches
the holding position, second coil spring 8b as well as first coil
spring 8a begins to compress and deform. Hence, immediately when
the oscillated angle of cam ring 5 increases a very small angle
from the holding position, the spring load rapidly discontinuously
increases up to a load W4 with a less change in the spring
displacement. The load W4 is equal to the summed value (W2+W3) of
the load W2 and an initial set load W3 of second coil spring
8b.
[0164] With cam ring 5 displaced at an angular position between the
holding position (FIG. 10) and the minimum-eccentricity position
(FIG. 11), first and second coil springs 8a-8b are both compressed
and deformed. The spring load becomes the summed value of the
spring load produced by first coil spring 8a and the spring load
produced by second coil spring 8b. Thus, the spring load tends to
increase at a gradient corresponding to the summed value of the
spring constant of first coil spring 8a and the spring constant of
second coil spring 8b, in proportion to the displacement of first
and second coil springs 8a-8b from the holding position. When cam
ring 5 reaches the minimum-eccentricity position (see FIG. 11), the
spring load becomes a load W5 (>W4), corresponding to the
displacements of first and second coil springs 8a-8b at that time.
Therefore, with cam ring 5 displaced at an angular position between
the holding position (FIG. 10) and the minimum-eccentricity
position (FIG. 11), the moment Ts, resulting from biasing member 8,
becomes identical to the sum (Ts1+Ts2) of the moment Ts1, produced
by the spring load of first coil spring 8a and the moment Ts2,
produced by the spring load of second coil spring 8b.
[0165] As discussed above, the spring displacement versus load
characteristic of biasing member 8 is designed as a nonlinear
characteristic, in which, the load (that is, the biasing force)
increases discontinuously, as the oscillated amount (the oscillated
angle) of cam ring 5 increases. That is, biasing member 8 has a
discontinuous spring characteristic that the spring load increases
rapidly discontinuously at the holding position over the entire
range of spring load, ranging from the initial setting position via
the holding position to the minimum-eccentricity position. The
spring constant of biasing member 8 becomes identical to the spring
constant of first coil spring 8a within a first range of
oscillating motion of cam ring 5, ranging from the initial setting
position to the holding position. The spring constant of biasing
member 8 becomes identical to the summed value of the spring
constant of first coil spring 8a and the spring constant of second
coil spring 8b within a second range of oscillating motion of cam
ring 5, ranging from the holding position to the
minimum-eccentricity position. The spring constant of biasing
member 8, i.e., the load (the biasing force) per unit spring
displacement tends to rapidly discontinuously increase.
[0166] The previously-discussed nonlinear characteristic is
obtained by the double spring structure, which is comprised of
first coil spring 8a, which permanently biases cam ring 5
counterclockwise regardless of the oscillated amount of cam ring 5,
and second coil spring 8b, which applies its biasing force (spring
force) to cam ring 5 only when the oscillated amount of cam ring 5
exceeds a predetermined amount. That is, biasing member 8 is
configured, so that cam ring 5 is forced by means of only one
spring (i.e., first coil spring 8a) when the oscillated amount of
cam ring 5 is small, and that cam ring 5 is forced by means of a
plurality of springs (i.e., first and second coil springs 8a-8b)
when the oscillated amount of cam ring 5 is large.
[0167] Referring now to FIG. 13, there are shown the engine-speed
versus pump-discharge-pressure characteristic curves. The
characteristic curve, indicated by the solid line (a) in FIG. 13,
shows the relationship between the engine speed (i.e., the
revolution speed of pump VP) and discharge pressure Pd, in the
variable displacement vane pump VP of the first embodiment. On the
other hand, the characteristic curve, indicated by the broken lines
(b)-(c) in FIG. 13, shows a typical engine-speed versus discharge
pressure characteristic, generally utilized by usual engines.
[0168] Hydraulic pressure, required for the engine, is determined
mainly by hydraulic pressure required for lubrication of bearings
of the engine crankshaft. Thus, as can be seen from the broken line
(c) of FIG. 13, discharge pressure Pd tends to increase, as the
engine speed increases. Generally, in the case of
hydraulically-operated VTC-equipped engines, pump discharge
pressure is also used as a working-fluid-pressure source for the
VTC system. In order to enhance the operation responsiveness of the
VTC system, a predetermined pressure level of discharge pressure Pd
(see a pressure level P.sub.1* indicated on the broken line (b) in
FIG. 13) is required from the point of time when the engine speed
is still low. As discussed previously, the usual VTC-equipped
engine has the engine-speed versus discharge pressure
characteristic, obtained by combining the broken lines (b)-(c) with
each other.
[0169] In contrast, in the first embodiment, by virtue of the
previously-noted nonlinear characteristic of biasing member 8, pump
VP exhibits the engine-speed versus discharge pressure
characteristic indicated by the solid line (a) in FIG. 13. The
relationship between discharge pressure Pd and engine speed (pump
revolution speed) is hereunder described in detail each and every
speed range, namely, a first speed range Ne.sub.1-2, a second speed
range Ne.sub.2-3, a third speed range Ne.sub.3-4, and a fourth
speed range Ne.sub.4-5.
[0170] In the first speed range Ne.sub.1-2 just after engine
starting, in which engine speed is still low, the counterclockwise
moment Ts1, resulting from the initial set load W1 of biasing
member 8 (first coil spring 8a) becomes greater than the clockwise
moment Ta-Tb, resulting from discharge pressure Pd of pump VP, and
thus cam ring 5 is kept at the initial setting position shown in
FIG. 5. With cam ring 5 kept at its initial setting position, the
eccentricity |OP| of geometric center "P" with respect to the axis
"O" becomes maximum, the pump discharge capacity becomes maximum,
and thus pump VP has a high discharge pressure rise characteristic
that discharge pressure Pd can be risen rapidly in accordance with
an engine speed rise (see a steep discharge pressure rise in the
first speed range Ne.sub.1-2 in FIG. 13).
[0171] As soon as discharge pressure Pd becomes greater than or
equal to a pressure level P2, the clockwise moment Ta-Tb, resulting
from discharge pressure Pd, becomes greater than the
counterclockwise moment Ts1, resulting from the initial set load W1
of biasing member 8 (first coil spring 8a), and thus cam ring 5
begins to oscillate in the direction that the eccentricity |OP| of
geometric center "P" with respect to the axis "O" reduces. In the
second speed range Ne.sub.2-3, discharge pressure Pd rises from the
pressure level P2 up to a pressure level P3, in accordance with an
engine speed rise. During the period in which discharge pressure Pd
is rising from the pressure level P2 up to the pressure level P3,
assuming that the clockwise moment Ta-Tb, resulting from discharge
pressure Pd, continuously exceeds the moment Ts1, resulting from
the load (ranging from W1 to W2) of biasing member 8 (first coil
spring 8a) compressed, cam ring 5 can be continuously oscillated in
the previously-noted direction that the eccentricity |OP| reduces.
During such a clockwise oscillating motion of cam ring 5 in the
second speed range Ne.sub.2-3, a discharge pressure buildup,
resulting from an engine speed rise, is canceled by a discharge
pressure reduction, resulting from a decrease in the pump discharge
capacity. For the reasons discussed above, the gradient of a
discharge pressure rise with respect to an engine speed rise,
produced in the second speed range Ne.sub.2-3, tends to be less
than that produced in the first speed range Ne.sub.1-2. Thus, in
the second speed range Ne.sub.2-3, pump VP has a slow discharge
pressure rise characteristic that discharge pressure Pd can be
risen slowly in accordance with an engine speed rise (see a slow
discharge pressure rise in the second speed range Ne.sub.2-3 in
FIG. 13).
[0172] When discharge pressure Pd reaches a pressure level P3, the
clockwise moment Ta-Tb, resulting from discharge pressure Pd,
becomes identical to the moment Ts1, resulting from the load W2 of
biasing member 8 (first coil spring 8a). In the third speed range
Ne.sub.3-4, discharge pressure Pd rises from the pressure level P3
up to a pressure level P4, in accordance with an engine speed rise.
During the period in which discharge pressure Pd is rising from the
pressure level P3 up to the pressure level P4, the clockwise moment
Ta-Tb, resulting from discharge pressure Pd, is balanced to the
counterclockwise moment Ts, resulting from the summed spring load
(ranging from W2 to W4) of first and second coil springs 8a-8b.
Thus, cam ring 5 remains kept at its holding position, with no
further oscillating motion (with no further clockwise rotation).
The rate of change of the volume of each of pump chambers r1-r7,
obtained in the holding position of cam ring 5, is lower than that
obtained in the initial setting position. Hence, the pump discharge
capacity, obtained in the third speed range Ne.sub.3-4, is less
than that obtained in the first speed range Ne.sub.1-2. Regarding
the pump discharge capacity, the third speed range Ne.sub.3-4,
differs from the second speed range Ne.sub.2-3. That is, the pump
discharge capacity tends to decrease in the second speed range
Ne.sub.2-3, whereas the pump discharge capacity remains unchanged
and becomes a fixed value in the third speed range Ne.sub.3-4. For
the reasons discussed above, the gradient of a discharge pressure
rise with respect to an engine speed rise, produced in the third
speed range Ne.sub.3-4, tends to be less than that produced in the
first speed range Ne.sub.1-2, and greater than that produced in the
second speed range Ne.sub.2-3. That is, in the third speed range
Ne.sub.3-4, pump VP has a moderate discharge pressure rise
characteristic that discharge pressure Pd can be risen moderately
in accordance with an engine speed rise (see a moderate discharge
pressure rise in the third speed range Ne.sub.3-4 in FIG. 13).
[0173] When, due to a further discharge pressure rise, discharge
pressure Pd becomes greater than or equal to the predetermined
pressure level P4, the clockwise moment Ta-Tb, resulting from
discharge pressure Pd, becomes greater than the counterclockwise
moment Ts, resulting from the spring load W4 (the summed spring
load of first and second coil springs 8a-8b). Cam ring 5 begins to
oscillate again in the direction that the eccentricity |OP| of
geometric center "P" with respect to the axis "O" reduces. Thus, in
the fourth speed range Ne.sub.4-5, discharge pressure Pd rises from
the pressure level P4 up to a pressure level P5, in accordance with
a further engine speed rise. During the period in which discharge
pressure Pd is rising from the pressure level P4 up to the pressure
level P5, assuming that the clockwise moment Ta-Tb, resulting from
discharge pressure Pd, continuously exceeds the moment Ts,
resulting from the load (ranging from W4 to W5) of biasing member 8
(first and second coil springs 8a-8b) compressed, cam ring 5 can be
continuously oscillated in the previously-noted direction that the
eccentricity |OP| reduces. Hence, in a similar manner to the second
speed range Ne.sub.2-3, the gradient of a discharge pressure rise
with respect to an engine speed rise, produced in the fourth speed
range Ne.sub.4-5, tends to be less than those produced in the first
and third speed ranges Ne.sub.1-2 and Ne.sub.3-4. Thus, in the
fourth speed range Ne.sub.4-5, pump VP has a slow discharge
pressure rise characteristic that discharge pressure Pd can be
risen slowly in accordance with an engine speed rise (see a slow
discharge pressure rise in the fourth speed range Ne.sub.4-5 in
FIG. 13).
[0174] The shape of the specific engine-speed versus discharge
pressure characteristic curve (indicated by the solid line (a) in
FIG. 13) of pump VP of the first embodiment, in which the high
discharge pressure rise characteristic of first speed range
Ne.sub.1-2, the slow discharge pressure rise characteristic of
second speed range Ne.sub.2-3, the moderate discharge pressure rise
characteristic of third speed range Ne.sub.3-4, and the slow
discharge pressure rise characteristic of fourth speed range
Ne.sub.4-5 are combined with each other, is, as a whole, analogous
to the shape of the engine-speed versus discharge pressure
characteristic curve, obtained by combining the broken lines
(b)-(c) with each other and generally required for usual
VTC-equipped engines.
[0175] (Operation of VTC)
[0176] The operation of the VTC system employing pump VP having the
specific engine-speed versus discharge pressure characteristic, as
indicated by the solid line (a) in FIG. 13, is hereunder described
in detail in reference to FIGS. 14A-14C, 15A-15C, and 16A-16C. In a
similar manner to FIG. 2, FIGS. 14A-16C, FIGS. 14A, 15A, and 16A
show the front elevation views, schematically illustrating angular
positions of vane member 300. FIGS. 14B, 15B, and 16B show the
cross sections, taken along the line J-J of FIG. 2, and
schematically illustrating positions of retractable lock piston
510. In a similar manner to FIG. 1, FIGS. 14C, 15C, and 16C show
the cross sections of directional control valve 450, cut along the
x-axis direction, and schematically illustrating axial positions of
valve spool 490. FIGS. 14A-14C show the relationship among the
angular position of vane member 300, the position of retractable
lock piston 510, and the axial position of valve spool 490, in an
engine stopped state. FIGS. 15A-15C show the relationship among the
angular position of vane member 300, the position of retractable
lock piston 510, and the axial position of valve spool 490, during
an engine startup period. FIGS. 16A-16C show the relationship among
the angular position of vane member 300, the position of
retractable lock piston 510, and the axial position of valve spool
490, at middle engine-speed operation. As previously described, the
phase converter (the variable valve actuation mechanism) is
comprised of vane member 300, phase-converter housing 102,
variable-volume phase-advance chambers 311, 321, 331, and 341,
variable-volume phase-retard chambers 312, 322, 332, and 342, and
hydraulic pressure supply-and-exhaust mechanism 400.
[0177] In the engine stopped state, pump VP does not yet come into
operation, and also there is no output of exciting current from
controller CU to electromagnetic coil 482 of solenoid 480 of
directional control valve 450. Thus, as seen in FIG. 14C, valve
spool 490 is forced in the positive x-axis direction by the spring
force of return spring RS, and held at the maximum leftward axial
position (the spring-loaded position). Hence, there is no supply of
working-fluid pressure to each of phase-advance chambers 311-341
and there is no supply of working-fluid pressure to each of
phase-retard chambers 312-342.
[0178] On the other hand, as seen in FIG. 14A, vane member 300
rotates relative to timing sprocket 100 (or phase-converter housing
102) in the counterclockwise direction opposite to the
timing-sprocket rotation direction (see the direction of rotation
of timing sprocket 100 indicated by the arrow in FIG. 2) by
alternating torque, occurring at camshaft 200 just after the engine
has stopped. As a result of this, vane member 300 can be positioned
at its maximum phase-retard position in advance. At this time, as
seen in FIG. 14B, tapered head portion 511 of lock piston 510 is
brought into engagement with engaging hole 521 of cup-shaped
engaging-hole structural member 520 of rear cover 104 by the spring
force of return spring 540 of lock mechanism 500, and thus free
rotation of vane member 300 relative to phase-converter housing 102
is prevented.
[0179] Next, when beginning to crank the engine by turning an
ignition key (not shown) ON, there is no output of control current
from controller CU to electromagnetic coil 482 for a brief moment
(several seconds) from the beginning of cranking. Thus, as seen in
FIG. 15C, valve spool 490 still remains kept at the maximum
leftward axial position (the spring-loaded position) by the spring
force of return spring RS. With valve spool 490, held at its
maximum leftward position, fluid communication between supply port
471 and second port 473 is established, and fluid communication
between second port 473 and second drain port 475 is blocked, and
fluid communication between first port 472 and first drain port 474
is established.
[0180] Thus, as indicated by the arrow in FIG. 15C, the fluid
pressure (discharge pressure Pd) of working fluid discharged from
pump VP flows from supply passage 430 through supply port 471 into
valve housing 470, and then flows via second port 473 into second
hydraulic line 420. Thereafter, the working-fluid pressure is
further delivered via branch passages 422-425 into each of
phase-retard chambers 312-342 (see FIG. 15A). The pressure in each
of phase-retard chambers 312-342 rises, as discharge pressure Pd of
pump VP builds up.
[0181] As seen in FIG. 15B, due to the previously-discussed
pressure rise in phase-retard chamber 342, the risen hydraulic
pressure in phase-retard chamber 342 is delivered via second oil
hole 344h into pressure-receiving chamber 550, and acts on the
pressure-receiving surface of flanged portion 513 of lock piston
510. When discharge pressure Pd of pump VP exceeds a predetermined
pressure value, the hydraulic pressure acting on the
pressure-receiving surface of flanged portion 513 becomes greater
than the spring force of return spring 540, and as a result lock
piston 510 begins to move in the positive x-axis direction. When
discharge pressure Pd becomes greater than or equal to a pressure
level P1, tapered head portion 511 is completely forced out of
engaging hole 521 of cup-shaped engaging-hole structural member 520
of rear cover 104. In this manner, the locked state of lock
mechanism 500 of vane member 300 (i.e., the locked state of the
valve timing) becomes released, thereby permitting free rotation of
vane member 300 so as to enable a phase change by means of the VTC
system.
[0182] The previously-noted pressure level P1, at which lock
mechanism 500 of the VTC device becomes unlocked, can be realized
in the first speed range Ne.sub.1-2 shown in FIG. 13. The pressure
level P1 is lower than the pressure level P2 at which cam ring 5
begins to oscillate clockwise against the spring force of first
coil spring 8a. The discharge pressure Pd, corresponding to
pressure level P1, can be realized two or three seconds later after
the ignition key has been turned ON. At the point of time when lock
mechanism 500 of the VTC device becomes unlocked, cam ring 5 of
pump VP remains kept in the initial setting position before
beginning to compress first coil spring 8a by protruding portion 54
of cam-ring arm portion 5d, and thus the pump discharge capacity
becomes maximum. Hence, at the time when lock mechanism 500 of the
VTC device becomes unlocked, pump VP has a high discharge pressure
rise characteristic that discharge pressure Pd can be risen rapidly
in accordance with an engine speed rise (see a steep discharge
pressure rise in the first speed range Ne.sub.1-2 in FIG. 13).
[0183] When discharge pressure Pd is greater than or equal to the
pressure level P1 and less than the pressure level P2 even after
lock mechanism 500 of the VTC device has been unlocked, in a
similar to the engine stopped state, vane member 300 is still
maintained at the maximum phase-retard position under a
comparatively low hydraulic pressure supplied into each of
phase-retard chambers 312-342, even after the engine has been
cranked and started (see FIGS. 15A-15B), thereby enhancing the
engine startability. At this time, air staying in each of
phase-retard chambers 312-342 is compressed by the hydraulic
pressure, and then the compressed air, together with the hydraulic
pressure, acts to force vane member 300 toward the maximum
phase-retard side.
[0184] When the engine operating range reaches a middle speed range
(e.g., the second speed range Ne.sub.2-3 shown in FIG. 13) after
cranking, discharge pressure Pd becomes greater than the pressure
level P2. At this time, controller CU outputs a control command
signal (an exciting current) to electromagnetic coil 482 of
solenoid 480 of directional control valve 450. Hence, valve spool
490 moves from the maximum leftward axial position (the
spring-loaded position shown in FIG. 15C) in the negative x-axis
direction, and then positioned in the maximum rightward axial
position shown in FIG. 16C. With valve spool 490, held at its
maximum rightward position, fluid communication between supply port
471 and first port 472 is established, and fluid communication
between second port 473 and second drain port 475 is established,
and fluid communication between first port 472 and first drain port
474 is blocked.
[0185] Thus, as indicated by the arrows in FIG. 16C, discharge
pressure Pd of pump VP flows from supply passage 430 through supply
port 471 into valve housing 470, and then flows via first port 472
into first flow-passage structure 411 of first hydraulic line 410.
Thereafter, discharge pressure Pd is further delivered via branch
passages 412-415 into each of phase-advance chambers 311-341 (see
FIG. 16A). Hence, the pressure in each of phase-advance chambers
311-341 becomes high, while the pressure in each of phase-retard
chambers 312-342 becomes low, since the working oil in each of
phase-retard chambers 312-342 is drained through second hydraulic
line 420 and second drain port 475 into oil pan 460.
[0186] As seen in FIG. 16B, although the hydraulic pressure in
pressure-receiving chamber 550 tends to drop, due to the
previously-discussed pressure rise in phase-advance chamber 341,
the risen hydraulic pressure in phase-advance chamber 341 is
delivered via first oil groove 343g into tapered head portion 511.
Thus, lock piston 510 receives the hydraulic pressure acting in the
positive x-axis direction, and whereby the unlocked state, where
lock piston 510 is forced out of engaging hole 521 against the
spring force of return spring 540, is maintained.
[0187] When the hydraulic pressure in each of phase-advance
chambers 311-341 is rising up due to discharge pressure Pd
developing above the pressure level P2, as shown in FIG. 16A, vane
member 300 rotates relative to timing sprocket 100 (or
phase-converter housing 102) from the angular position shown in
FIG. 15A in the same direction (i.e., in the clockwise direction)
as the timing-sprocket rotation direction (see the direction of
rotation of timing sprocket 100 indicated by the arrow in FIG. 2).
As a result, the angular phase of camshaft 200 relative to the
crankshaft can be rapidly varied to the phase-advance side, and
thus a valve overlapping period during which the intake and exhaust
valves are both open, becomes slightly great.
[0188] In contrast, when, in the second speed range Ne.sub.2-3, the
hydraulic pressure in each of phase-advance chambers 311-341 drops
due to some kind of factors, such as an engine speed drop, the
angular phase of camshaft 200 relative to the crankshaft can be
rapidly returned to the phase-retard side, and thus the valve
overlapping period becomes decreased. At this time, discharge
pressure Pd remains kept at a pressure level higher than the
pressure level P2, lock mechanism 500 remains kept at the unlocked
state.
[0189] As discussed above, in the middle engine speed range, that
is, in the second speed range Ne.sub.2-3, discharge pressure Pd is
greater than or equal to the pressure level P2 and less than the
pressure level P3, cam ring 5 can oscillate or rotate, while being
forced counterclockwise by only one spring, namely, only the spring
force of first coil spring 8a. Hence, in the middle engine speed
range, pump VP has a slow discharge pressure rise characteristic
that discharge pressure Pd can be risen slowly in accordance with
an engine speed rise (see a slow discharge pressure rise in the
second speed range Ne.sub.2-3 in FIG. 13). The discharge pressure
level P2, at which cam ring 5 begins to oscillate clockwise against
only the spring force of first coil spring 8a, is slightly higher
than the pressure level P.sub.1*, indicated on the broken line (b)
in FIG. 13 and required for the enhanced operation responsiveness
of the VTC system in a low engine speed range. Thus, in the second
speed range Ne.sub.2-3, free rotation of vane member 300 relative
to timing sprocket 100 (phase-converter housing 102) can be
permitted, while maintaining the unlocked state of lock mechanism
500. In other words, even in a biasing state where cam ring 5 is
forced counterclockwise by means of only one spring (first coil
spring 8a), a hydraulic pressure, needed to operate the VTC device,
can be always produced during operation of pump VP.
[0190] When the operating range of the engine exceeds the middle
speed range (e.g., the second speed range Ne.sub.2-3 shown in FIG.
13), and enters the third speed range Ne.sub.3-4, discharge
pressure Pd becomes greater than or equal to the pressure level P3.
The output of the control command signal (exciting current) from
controller CU to electromagnetic coil 482 of solenoid 480 is
maintained, and thus high hydraulic pressure can be continuously
supplied into each of phase-advance chambers 311-341. Hence, vane
member 300 further rotates relative to timing sprocket 100 in the
clockwise direction, and thus the angular phase of camshaft 200
relative to the crankshaft further varies to the phase-advance
side. Finally, vane member 300 is held at the maximum phase-advance
position at which the volumetric capacity of each of phase-advance
chambers 311-341 becomes a maximum, and thus the valve overlapping
period becomes a maximum.
[0191] In the third speed range Ne.sub.3-4, discharge pressure Pd
is greater than or equal to the pressure level P3 and less than the
pressure level P4, cam ring 5 is kept at the holding position.
Hence, pump VP has a moderate discharge pressure rise
characteristic that discharge pressure Pd can be risen moderately
in accordance with an engine speed rise (see a moderate discharge
pressure rise in the third speed range Ne.sub.3-4 in FIG. 13), in
other words, a discharge pressure rise characteristic that
discharge pressure Pd can be risen moderately at a slightly greater
gradient of a discharge pressure rise with respect to an engine
speed rise as compared to that of an engine-speed versus discharge
pressure characteristic (see the characteristic curve, indicated by
the broken line (c) in FIG. 13), required for lubrication of moving
and/or sliding engine parts, such as engine crankshaft
bearings.
[0192] When the engine operating range reaches a high speed range
(e.g., the fourth speed range Ne.sub.4-5 shown in FIG. 13),
discharge pressure Pd becomes greater than the pressure level P4.
In the high speed range, cam ring 5 can oscillate or rotate, while
being forced counterclockwise by two springs, namely, the summed
spring force of first and second coil springs 8a-8b. Hence, in a
similar manner to the second speed range Ne.sub.2-3 (the middle
speed range), in the fourth speed range Ne.sub.4-5 (in the high
speed range), pump VP has a slow discharge pressure rise
characteristic that discharge pressure Pd can be risen slowly in
accordance with an engine speed rise (see a slow discharge pressure
rise in the fourth speed range Ne.sub.4-5 in FIG. 13). A pressure
level P5 of discharge pressure Pd, produced when the engine speed
becomes maximum, becomes slightly higher than a pressure level
P.sub.2*, indicated on the broken line (c) in FIG. 13 and required
for lubrication of moving and/or sliding engine parts at the
maximum engine speed.
COMPARATIVE EXAMPLE
[0193] The operation and effects of the first embodiment are
hereunder explained, while comparing with the comparative example.
The comparative example has a variable displacement vane pump
construction differing from pump VP of the first embodiment. For
the same applicable condition, that is, on the assumption that the
variable displacement vane pump of the comparative example is
applied to the same VTC system to which pump VP of the first
embodiment is applied, the engine-speed versus discharge pressure
characteristic (see the characteristic curve (d) indicated by the
solid line in FIG. 17), obtained by the comparative example,
differs from the characteristic curve (see the characteristic curve
(a) indicated by the solid line in FIG. 13), obtained by the first
embodiment. In more detail, the vane pump of the comparative
example differs from pump VP of the first embodiment, only in the
following respects.
[0194] That is, the vane pump of the comparative example has a
control oil chamber defined between the inner periphery of a
vane-pump housing and the outer periphery of a cam ring and
partitioned by means of a seal member. Concretely, the seal member
is provided at the position substantially symmetrical to the
fulcrum (a pivot pin) of oscillating motion of the cam ring with
respect to the axis of a vane rotor. The internal space, which is
defined between the cam-ring outer peripheral surface and the
pump-housing inner peripheral surface, is partitioned into two
spaces by means of both the seal member and the fulcrum (the pivot
pin). These two spaces are partitioned from each other in a
fluid-tight fashion.
[0195] One space of the two partitioned spaces serves as a control
oil chamber. Also provided is a hydraulic pressure control
mechanism (including working-fluid flow control valves and the
like) formed integral with the pump housing as a pump unit or
provided separately from the pump, for controlling the hydraulic
pressure in the control oil chamber. The pump housing is formed
with a plurality of hydraulic ports, connected to the hydraulic
pressure control mechanism, so as to supply and exhaust working oil
(hydraulic pressure) to and from the control oil chamber. By the
pressure, resulting from the hydraulic pressure delivered into the
control oil chamber, and acting on a specific area (one partitioned
space) of the cam-ring outer peripheral surface, corresponding to
the circumferentially extending control oil chamber, the cam ring
can oscillate about its fulcrum, so as to vary the pump discharge
capacity. A biasing member (such as a spring) is provided in the
other partitioned space, for forcing the cam ring in one direction
(toward a spring-loaded position) against the pressure, resulting
from the hydraulic pressure delivered into the control oil chamber
defined on the cam-ring outer periphery, and acting on the specific
area (the one partitioned space). In the case of the vane pump of
the comparative example, regarding the pressure-receiving area of
the cam-ring inner peripheral surface on which discharge pressure
from the discharge port acts and which is divided into two
segmented pressure-receiving areas with respect to the fulcrum of
oscillating motion of the cam ring, these two segmented areas
accord with each other or do not accord with each other, depending
on oscillating motion of the cam ring. This is because the
comparative example is based on a prerequisite that oscillating
motion of the cam ring can be controlled by a control hydraulic
pressure acting on the cam-ring outer peripheral surface. Notice
that the comparative example never takes into account a control of
oscillating motion of the cam ring by the hydraulic pressure acting
on the cam-ring inner peripheral surface.
[0196] First, the comparative example requires the control oil
chamber defined on the cam-ring outer periphery, and thus the
comparative example has the difficulty of reducing the number of
component parts constructing the vane pump. Concretely the
previously-discussed seal member must be installed to define the
control oil chamber. Additionally, the machining accuracy of each
of pump component parts, to which the seal member is fitted, must
be enhanced. This leads to the problem of increased vane pump
manufacturing costs. Furthermore, the comparative example requires
the previously-discussed plural hydraulic ports and hydraulic
pressure control mechanism, for controlling or adjusting the
hydraulic pressure in the control oil chamber. The comparative
example has the difficulty of compactly designing the vane pump. In
addition to the above, the comparative example requires the precise
setting (fine adjustment) of an interference fit between mating
parts after assembling the pump unit. Such a seal member is very
hard to install or assemble accurately. For the reasons discussed
above, the comparative example has several drawbacks, that is,
increased oil leakages and contamination due to increased fittings,
increased system installation time and costs, and increased service
time.
[0197] As discussed above, in the case of the comparative example,
hydraulic pressure in the control oil chamber acts between the
cam-ring outer peripheral surface and the pump-housing inner
peripheral surface. Thus, it is necessary to prevent undesirable
oil leakage from the interior space of the pump housing to the
exterior space. From the viewpoint of reduced leakage, the
thickness of the flanged portion of the pump housing has to be
increased and a gasket has to be used to enhance a fluid-tight
performance of the vane pump. It is difficult to compactly design
the pump, and also it is difficult to reduce the pump system
costs.
[0198] In contrast to the above, in the case of pump VP of the
first embodiment, the moment of force about the fulcrum "Q" of
oscillating motion of cam ring 5 results from the discharge
pressure acting on cam-ring inner peripheral surface 50, so as to
adjust or control the eccentricity |OP| of geometric center "P" of
cam-ring inner peripheral surface 50 with respect to the axis "O"
of drive shaft 3 (or vane rotor 4). Hence, this eliminates the need
for the control oil chamber arranged on the cam-ring outer
periphery. Therefore, it is possible to reduce the number of seal
members, such as oil seals and gaskets, thus realizing reduced
number of pump component parts. Additionally, it is possible to
solve the problem of the higher machining accuracy of each of pump
component parts, to which the seal member is fitted, thus further
reducing pump manufacturing costs. Furthermore, it is possible to
eliminate the necessity of the plural hydraulic ports and hydraulic
pressure control mechanism, required for controlling the hydraulic
pressure in the control oil chamber. This contributes to the
compact vane pump system and smaller space requirements of overall
pump system. Additionally, the installation process of the seal
member, defining the control oil chamber between the cam-ring outer
periphery and the pump-housing inner periphery, is unnecessary.
This contributes to the lower system installation time and costs,
and lower service time.
[0199] In addition to the above, in the first embodiment, any
control hydraulic pressure cannot be supplied into clearance space
CL, defined between an outer peripheral surface 50a of cam ring 5
and inner peripheral surface 13a of pump-housing peripheral wall
13, and hence the pressure in clearance space CL can be held at a
low pressure level (approximately, an atmospheric pressure level).
Thus, it is unnecessary to increase the thickness of flanged
portion 14 of pump housing 1, and also it is possible to eliminate
the necessity of a gasket used to enhance a fluid-tight performance
of the vane pump. This contributes to the compactly-designed pump
VP, smaller space requirements of overall pump system, and lower
pump system costs.
[0200] Moreover, according to the vane pump system configuration of
the first embodiment, cam ring 5 is pivoted by means of pivot pin 9
(serving as the fulcrum "Q" of oscillating motion of cam ring 5).
Hence, biasing member 8 can be laid out or installed at an
arbitrary circumferential position around the entire circumference
of cam ring 5, at which cam ring 5 can be forced in the direction
that the eccentricity |OP| of geometric center "P" of cam-ring
inner peripheral surface 50 with respect to the axis "O" of drive
shaft 3 (or vane rotor 4) increases. This contributes to the
increased layout flexibility of biasing member 8. Therefore, it is
possible to optimize the layout of biasing member 8 relative to
inlet hole 16a and inlet port 16b, thus effectively enhancing the
pump efficiency in particular on the inlet side (described later in
reference to the second to fourth embodiments).
[0201] The force, which produces oscillating motion of cam ring 5,
is determined depending on discharge pressure Pd acting on cam-ring
inner peripheral surface 50 and the position of the fulcrum "Q" of
oscillating motion of cam ring 5. Hence, as the position of the
fulcrum "Q" of oscillating motion of cam ring 5 approaches closer
to the middle position at which the clockwise moment Ta about the
fulcrum "Q" and the counterclockwise moment Tb about the fulcrum
"Q" are balanced to each other, the cam-ring oscillating force,
resulting from discharge pressure Pd acting on cam-ring inner
peripheral surface 50, can be set to a smaller value. As a result,
it is possible to set the spring force, produced by biasing member
8 against the cam-ring oscillating force resulting from discharge
pressure Pd, to a smaller value. This contributes to the downsized
biasing member 8, smaller installation space requirement of biasing
member 8 (downsized spring chamber 15), and increased layout
flexibility of biasing member 8, thus realizing the compact vane
pump system and smaller space requirements of overall pump
system.
[0202] Secondly, the comparative example does not employ a biasing
member of a nonlinear characteristic. Therefore, in the comparative
example, it is difficult to reconcile the enhanced operation
responsiveness of the VTC system and a reduction in engine power
loss, which loss is caused by the vane pump employing a biasing
member of a linear characteristic. Generally, the vane pump can
output a predetermined pressure level of discharge pressure Pd from
a low pump revolution speed range (a low engine speed range). As
can be seen from the first transition of the engine-speed versus
discharge pressure characteristic curve (d) indicated by the solid
line in FIG. 17, in the comparative example, the eccentricity |OP|
of the geometric center of the cam-ring inner peripheral surface
with respect to the axis of the vane rotor is set, so that the
gradient of a discharge pressure rise with respect to an engine
speed rise becomes a predetermined gradient ".alpha." at the
initial setting state, while utilizing the previously-discussed
general discharge pressure characteristic. By virtue of such a
steep gradient ".alpha.", the vane pump system of the comparative
example ensures or achieves the pressure level P.sub.1*, required
for operating the VTC system from an engine low speed range.
[0203] Assuming that the steep gradient ".alpha." remains
unchanged, the difference (the deviation) between the minimum
required hydraulic pressure (see the characteristic curve,
indicated by the broken lines (b)-(c) in FIG. 17) and discharge
pressure Pd (see the characteristic curve (d) indicated by the
solid line in FIG. 17) tends to increase, as the engine speed
increases. To avoid this, when discharge pressure Pd rises up to a
pressure level P.sub.3*, indicated on the solid line (d) in FIG. 17
and identical to the maximum required pressure level P.sub.2*
indicated on the broken line (c) in FIG. 13 and required for
lubrication of moving and/or sliding engine parts, oscillating
motion of the cam ring is carried out, so as to reduce the pump
discharge capacity. In this manner, the pump system of the
comparative example can suppress discharge pressure Pd from
unnecessarily rising, thereby suppressing an increase in engine
power loss. However, in the case of the comparative example, the
pump discharge pressure characteristic indicated by the solid line
(d) in FIG. 17 remarkably deviates from the actually-required pump
discharge pressure characteristic indicated by the broken lines
(b)-(c) in FIG. 17. Hence, as can be seen from the left-hand
diagonal shading area between the characteristic curve (d) and the
characteristic curve (b)-(c) in FIG. 17, in the pump system of the
comparative example, there is a great engine power loss (a great
energy waste) corresponding to the shading area.
[0204] To reduce the power loss, suppose that the pump discharge
pressure characteristic is changed from the characteristic
indicated by the solid line (d) in FIG. 17 to a characteristic (e)
indicated by the one-dotted line in FIG. 17. In such a case,
oscillating motion of the cam ring is carried out at an earlier
timing, as soon as discharge pressure Pd rises up to a lower
pressure level (see the intersection point between the
characteristic curve indicated by the solid line (d) and the
characteristic curve (e) indicated by the one-dotted line in FIG.
17) as compared to the pressure level P.sub.3* indicated on the
solid line (d) in FIG. 17. This contributes to a reduction in power
loss. However, there is a limit to the reduced power loss.
[0205] In contrast, pump VP of the first embodiment employs biasing
member 8 of the nonlinear characteristic as previously described.
The discharge pressure characteristic of pump VP (see the
engine-speed versus discharge pressure characteristic curve
indicated by the solid line (a) in FIG. 13) can be approximated to
the minimum required hydraulic pressure characteristic curve
indicated by the broken lines (b)-(c) in FIGS. 13 and 17, as much
as possible. Hence, according to the vane pump system of the first
embodiment, it is possible to effectively reduce the power loss
(the energy waste), caused by an unnecessary discharge pressure
rise, while ensuring the enhanced operation responsiveness of the
VTC system.
[0206] On the one hand, the vane pump system having a discharge
pressure characteristic of a higher gradient of a discharge
pressure rise with respect to an engine speed rise in a low engine
speed range, is superior in the enhanced operation responsiveness
of the VTC system. This is because such a pump system can quickly
supply working-oil pressure to the VTC system even during an engine
startup period. On the other hand, the vane pump system having an
excessively high gradient of a discharge pressure rise with respect
to an engine speed rise in a low engine speed range, is inferior in
reduced power loss (reduced energy waste), in particular in a high
engine speed range.
[0207] As discussed previously, the vane pump system of the first
embodiment employs biasing member 8 of the nonlinear
characteristic, and therefore the discharge pressure characteristic
of pump VP (see the characteristic curve (a) in FIG. 13) can be
approximated to the minimum required hydraulic pressure
characteristic curve (see the characteristic curve (b)-(c) in FIGS.
13 and 17), as much as possible. For the reasons discussed above,
even when the gradient of a discharge pressure rise with respect to
an engine speed rise is set to a great gradient ".beta." in a low
engine speed range (in the first speed range Ne.sub.1-2), there is
a less risk increasing the power loss. Hence, in pump VP of the
first embodiment, it is possible to more remarkably shorten the
time interval from the engine starting point to the time when the
supply of working-oil pressure to the VTC system begins to occur,
by setting the gradient of a discharge pressure rise with respect
to an engine speed rise to the gradient ".beta." greater than the
gradient ".alpha." of the discharge pressure characteristic (d) of
the pump of the comparative example in the low speed range. As a
matter of course, this contributes to the remarkably enhanced
operation responsiveness of the VTC system. That is, by virtue of a
better initial discharge pressure buildup in the first speed range
Ne.sub.1-2, just after starting the engine, lock mechanism 500 of
the VTC device can be quickly shifted to its unlocked state, and
additionally the angular phase of camshaft 200 relative to the
crankshaft can be quickly varied to the phase-retard side.
[0208] As set forth above, the first embodiment enables a more
compact and light-weight variable displacement vane pump
construction, thereby allowing excellent mountability, and also
ensuring a superior pump efficiency, while simplifying the overall
pump system.
Effects of First Embodiment
[0209] Hereunder enumerated in detail are the effects of pump VP of
the first embodiment.
[0210] (1) Variable displacement vane pump VP of the first
embodiment includes rotor 4 driven by an internal combustion
engine, cam ring 5 configured to accommodate therein rotor 4 and
further configured to oscillate about a fulcrum "Q" of oscillating
motion along two axially opposed sidewalls (i.e., bottom face 10a
of basal portion 10 of pump housing 1 and bottom face 20a of
main-body portion 20 of pump cover 2) facing both sides of cam ring
5 respectively, a plurality of vanes 6a-6g, each of which is fitted
into rotor 4 to slide from rotor 4 toward inner peripheral surface
50 of cam ring 5 and set to be kept in abutted-engagement with
cam-ring inner peripheral surface 50, the vanes being configured to
define a plurality of working chambers (pump chambers r1-r7) in
cooperation with outer peripheral surface 42a of rotor 4, inner
peripheral surface 50 of cam ring 5, and the two axially opposed
sidewalls, biasing member 8 configured to force cam ring 5 in a
direction that the geometric center "P" of cam-ring inner
peripheral surface 50 and the rotation center "O" of rotor 4 are
spaced apart from each other, and an inlet portion (inlet port 16b,
inlet hole 16a) and a discharge portion (discharge port 17b,
discharge hole 17a) both formed in at least one of the two axially
opposed sidewalls, the inlet portion being configured to open into
a first group of working chambers (pump chambers r1, r2, r3, r4) of
the plurality of working chambers so as to extend over the first
group of working chambers (pump chambers r1, r2, r3, r4) within an
area where volumes of the first group of working chambers increase
(during rotary motion of rotor 4), and the discharge portion being
configured to open into a second group of working chambers (pump
chambers r4, r5, r6, r7) of the plurality of working chambers so as
to extend over the second group of working chambers (pump chambers
r4, r5, r6, r7) within an area where volumes of the second group of
working chambers decrease (during rotary motion of rotor 4). A
force, by which cam ring 5 can be oscillated against biasing member
8 in accordance with a buildup of a pressure in the discharge
portion, acts on cam-ring inner peripheral surface 50.
[0211] Thus, it is unnecessary to provide a control oil chamber and
a seal member on the side of the outer periphery of cam ring 5,
thereby realizing reduced number of pump component parts.
[0212] (2) Concretely, the fulcrum "Q" of oscillating motion of cam
ring 5 is laid out to be offset in a biasing direction of biasing
member 8 within an opening range of the discharge portion
(discharge port 17b).
[0213] Thus, during operation of pump VP, an area of cam-ring inner
peripheral surface 50, on which the pressure in the discharge
portion acts and which is segmented as a second pressure-receiving
area S2 extending in the biasing direction of biasing member 8 with
respect to the fulcrum "Q" (serving as a boundary) becomes
permanently smaller than an area of cam-ring inner peripheral
surface 50, on which the pressure in the discharge portion acts and
which is segmented as a first pressure-receiving area S1 extending
in a direction opposite to the biasing direction of biasing member
8 with respect to the fulcrum "Q". Hence, pump VP of the first
embodiment permits the force, by which cam ring 5 can be oscillated
against biasing member 8 in accordance with a buildup of the
pressure in the discharge portion, to be applied to cam ring 5 from
the side of cam-ring inner peripheral surface 50.
[0214] (3) More concretely, assuming that a point that a straight
line segment PQ, which links the fulcrum "Q" and the geometric
center "P" of cam-ring inner peripheral surface 50, intersects with
cam-ring inner peripheral surface 50 within the opening range of
the discharge portion (discharge port 17b), is defined as an
intersection point "R", the intersection point "R" is laid out to
be offset toward the circumferential end "C" of
beginning-of-discharge-port with respect to a center position (a
midpoint between two circumferential ends "C" and "D" of
discharge-port circular-arc shaped groove 17c) of the opening range
of the discharge portion (discharge port 17b). Biasing member 8 is
provided to force cam ring 5 so as to rotate cam ring 5 about the
fulcrum "Q" in a direction of the offset of intersection point "R",
which is offset from the center position.
[0215] That is to say, irrespective of whether cam ring 5 is
oscillating, there is a less change in the overlapping area that
cam-ring inner peripheral surface 50 overlaps with the discharge
portion (discharge port 17b), in other words, there is a less
change in the pressure-receiving area of cam-ring inner peripheral
surface 50 that receives discharge pressure Pd. Thus, the position
of a midpoint "S" of the above-mentioned pressure-receiving area
(circular-arc segment C'RD') can be approximated to the
intermediate position between the circumferential end "C" of
beginning-of-discharge-port and the circumferential end "D" of
end-of-discharge-port (of discharge-port circular-arc shaped groove
17c). Therefore, when the intersection point "R" is laid out to be
offset from the intermediate position, (i) a force Fb produced by
discharge pressure Pd received by second pressure-receiving area S2
extending in the biasing direction of biasing member 8 with respect
to the fulcrum "Q" and (ii) a force Fa produced by discharge
pressure Pd received by first pressure-receiving area S1 extending
in the direction opposite to the biasing direction with respect to
the fulcrum "Q" are put out of balance. Due to the unbalanced
forces Fa and Fb, a moment Ta-Tb of force about the fulcrum "Q" is
produced. On the other hand, biasing member 8 is configured to
produce a moment Ts about the fulcrum "Q" against the moment Ta-Tb,
by forcing cam ring 5 in the direction of the offset of the
intersection point "R".
[0216] In the first embodiment, the fulcrum "Q" is laid out on the
side of the discharge portion (discharge port 17b). In lieu
thereof, the fulcrum "Q" may be laid out on the side of the inlet
portion (inlet port 16b). For instance, suppose that the locations
of the discharge portion and the inlet portion are replaced with
each other without changing the position of the fulcrum "Q" as
shown in FIG. 5, and the pump is rotated in a reverse-rotational
direction. If the position of the intersection point "R" is set as
previously discussed even in such a layout of fulcrum "Q" on the
inlet-port side, this pump system can provide the same operation
and effects as the first embodiment. In a similar to the first
embodiment, in the case of the layout of fulcrum "Q" on the
inlet-port side, the fulcrum "Q" of oscillating motion of cam ring
5 is also laid out to be offset toward the circumferential end "C"
of beginning-of-discharge-port with respect to the center position
of the opening range of the discharge portion (discharge port
17b).
[0217] In the case of the layout of fulcrum "Q" on the inlet-port
side as previously discussed, the distance between the discharge
portion and fulcrum "Q" lengthens. For the same discharge pressure,
this fulcrum layout has a merit that, as compared to the first
embodiment, a greater moment Ta-Tb can be produced. In contrast, in
the first embodiment, the fulcrum "Q" of oscillating motion of cam
ring 5 is laid out on the discharge-port side, and thus the
distance between the discharge portion and fulcrum "Q" shortens.
Hence, the fulcrum layout of the first embodiment has a merit that
a produced moment Ta-Tb can be finely adjusted, thus facilitating
the suitable setting of the fulcrum "Q" of oscillating motion of
cam ring 5 and biasing member 8, and ensuring stable oscillating
action of cam ring 5.
[0218] (4) The fulcrum "Q" of oscillating motion of cam ring 5 is
laid out such that the integral .intg.S2dt of second segmented
pressure-receiving area S2 of cam-ring inner peripheral surface 50,
extending in the biasing direction of biasing member 8 with respect
to the fulcrum "Q", for a given cycle T, is less than the integral
.intg.S1dt of first segmented pressure-receiving area S1 of
cam-ring inner peripheral surface 50, extending in the direction
opposite to the biasing direction of biasing member 8 with respect
to the fulcrum "Q", for the given cycle T, that is,
.intg.S2dt<.intg.S1dt}.
[0219] That is to say, on the average, the relationship between the
magnitudes of first and second segmented pressure-receiving areas
S1 and S2 can be given by the inequality S1>S2 for the given
period T. Hence, during operation of pump VP, on the average, the
relationship between the magnitude of the force Fa resulting from
the fluid pressure acting on first segmented pressure-receiving
area S1 of cam-ring inner peripheral surface 50 and the magnitude
of the force Fb resulting from the fluid pressure acting on second
segmented pressure-receiving area S2 of cam-ring inner peripheral
surface 50 can be given by the inequality defined by Fa>Fb
during operation of pump VP. Therefore, it is possible to
continuously produce a moment-of-force Ta-Tb (>0), which causes
an oscillating or rotating motion of cam ring 5 in the direction
that the eccentricity |OP| of geometric center "P" of cam-ring
inner peripheral surface 50 with respect to the axis "O" of drive
shaft 3 (or the rotation center of rotor 4) reduces, during
operation of pump VP.
[0220] (5) Pump VP is configured such that a pressure, acting on
outer peripheral surface 50a of cam ring 5, is lower than the
pressure in the discharge portion. Hence, it is unnecessary to
tightly seal the inside and outside of pump housing 1. This
contributes to the compact vane pump system and lower pump system
costs.
[0221] (6) Concretely, atmospheric pressure is applied on outer
peripheral surface 50a of cam ring 5. Hence, the pressure
difference between the inside pressure and the outside pressure of
pump housing 1 can be set to approximately zero. Thus, it is
possible to further enhance the effect of the above-mentioned item
(5).
[0222] (7) An approximately uniform pressure is applied around an
entire circumference of cam-ring outer peripheral surface 50a.
Hence, pump VP of the first embodiment permits the force, by which
cam ring 5 can be oscillated against biasing member 8, to be
applied to cam ring 5 from only the side of cam-ring inner
peripheral surface 50, thereby ensuring a stable operating
characteristic of pump VP (a stable oscillating characteristic of
cam ring 5).
[0223] (8) The discharge portion includes grooves (circular-arc
shaped groove 17c, sector groove 17d, sector groove 23d) formed in
the two axially opposed sidewalls (i.e., bottom face 10a of basal
portion 10 of pump housing 1 and bottom face 20a of main-body
portion 20 of pump cover 2). Cam ring 5 has communication hole 51,
which is formed in the cam ring so as to axially penetrate cam ring
5 and through which the discharge portions (sector groove 17d of
the pump-housing side discharge port 17b, and sector groove 23d of
the pump-cover side discharge port 23) formed in the respective
sidewalls are communicated with each other. Working fluid is
discharged through a grooved portion (sector groove 17d) of at
least one of the discharge portions, configured to be substantially
conformable to a shape of communication hole 51 of cam ring 5, via
discharge hole 17a to an exterior space.
[0224] Therefore, the number of oil passages delivering working
fluid to discharge hole 17a can be increased, thus effectively
increasing a discharge of pump VP, and enhancing a discharging
ability of pump VP. Additionally, there is a less risk of
contaminant and debris accumulated in pump VP. Fluid pressures
applied on both sides of cam ring 5 from the discharge portions
formed in the respective sidewalls are almost balanced to each
other. Hence, a frictional force created between cam ring 5 and one
(pump housing 1) of the sidewalls and a frictional force created
between cam ring 5 and the other sidewall (pump cover 2) can be
both reduced or minimized, thus effectively reducing a force
required for oscillating motion of cam ring 5. Therefore, it is
possible to remarkably stabilize the pump operating characteristic,
while enhancing the durability of pump component parts.
[0225] (9) A fluid-flow passage cross-sectional area of
communication hole 51 is dimensioned to be greater than or equal to
that of discharge hole 17a. Hence, it is possible to reduce the
flow resistance to working-oil flow, caused by communication hole
51. Thus, it is possible to further enhance the effects of the
above-mentioned item (8).
[0226] (10) Communication hole 51 is formed into a circular-arc
shape (or a sector form) whose center is the fulcrum "Q" of
oscillating motion of cam ring 5. Thus, there is no risk that fluid
communication between the discharge portions (sector groove 17d of
discharge port 17b and sector groove 23d of discharge port 23) of
the sidewalls via communication hole 51 is blocked, even in the
presence of oscillating motion of cam ring 5. Additionally, even
when cam ring 5 is oscillating, there is a less change in the
overlapping area (the fluid-flow passage cross-sectional area)
between communication hole 51 and each of the discharge ports
(sector groove 17d of discharge port 17b and sector groove 23d of
discharge port 23). Thus, it is possible to stably provide the
operation and effects of the above-mentioned item (8).
[0227] (11) Communication hole 51 is configured to displace without
any change in an opening area of communication hole 51 opening into
discharge hole 17a, during oscillating motion of cam ring 5. That
is, during oscillating motion of cam ring 5, there is no change in
fluid-flow passage cross-sectional area of a working-fluid flow
passage oriented from discharge port 23 (sector groove 23d) of pump
cover 2 toward discharge hole 17a pump housing 1. Hence, there is a
less change in the fluid pressure applied from the side of
pump-cover discharge port 23 to cam ring 5. Thus, it is possible to
stably provide the operation and effects of the above-mentioned
item (8).
[0228] (12) A radial wall thickness L2 of a part of cam ring 5,
overlapping with the inlet portion and the discharge portion, is
dimensioned to be greater than a radial wall thickness L1 of the
other part of cam ring 5. That is, during operation of pump VP, the
fluid pressure in the inlet portion tends to become negative, and
thus the fluid pressure in the inlet portion tends to become lower
than the pressure acting on outer peripheral surface 50a of cam
ring 5 (or the pressure in clearance space CL). On the other hand,
the fluid pressure in the discharge portion tends to become higher
than the pressure acting on outer peripheral surface 50a of cam
ring 5 (or the pressure in clearance space CL). An internal space,
defined between cam ring 5 and each of the sidewalls in the axial
direction of pump VP, is a very small clearance space. Hence, this
design that the part of cam ring 5, overlapping with the inlet
portion and the discharge portion, is configured as a comparatively
thick-walled part contains lapped metal-to-metal sealing surfaces,
which form a virtually leak-proof seal. By the provision of the
thick-walled part of cam ring 5, it is possible to prevent leakage
of working fluid (working oil) from the discharge portion toward
the outer periphery of cam ring 5, and leakage of working fluid
from the outer periphery of cam ring 5 toward the inlet portion,
thereby enhancing the sealing performance of both of the inlet
portion and the discharge portion, and consequently enhancing the
pump efficiency.
[0229] Concretely, a radial width (a radial wall thickness) L2 of
cam-ring cylindrical portion 5a, overlapping with the
beginning-of-drawing-in-action region (see a part of cam-ring
cylindrical portion 5a near pump chambers r1, r2 in FIG. 5) of the
inlet portion (inlet port 16b) and the end-of-discharging-action
region (see a part of cam-ring cylindrical portion 5a near pump
chambers r6, r7 in FIG. 5) of the discharge portion (discharge port
17b), is dimensioned to be longer than a radial width (a radial
wall thickness) L1 of the other part of cam ring 5.
[0230] Therefore, even in the initial setting state where the
eccentricity |OP| of geometric center "P" of cam-ring inner
peripheral surface 50 with respect to the rotation center "O" of
vane rotor 4 becomes maximum, the enhanced sealing performance of
each of the inlet portion and the discharge portion can be ensured
sufficiently. That is, when the fluid pressure (discharge pressure
Pd) in the discharge portion is still insufficient at low engine
speeds (at low revolution speeds of pump VP), cam ring 5 becomes
kept in the initial setting position and thus the rate of change of
the volume of each of pump chambers r1-r7 becomes highest. Hence,
the fluid pressure in the beginning-of-drawing-in-action region of
the inlet portion (inlet port 16b) becomes low, while the fluid
pressure in the end-of-discharging-action region of the discharge
portion (discharge port 17b) becomes high. For the reasons
discussed above, it is necessary to enhance the sealing performance
between the inner and outer peripheries of cam ring 5, in
particular, at the beginning-of-drawing-in-action region of the
inlet portion (inlet port 16b) and at the end-of-discharging-action
region of the discharge portion (discharge port 17b). In the first
embodiment, by the provision of the thick-walled part of cam ring
5, even in the initial setting state (in an engine low speed
range), a proper radial distance L3 between cam-ring outer
peripheral surface 50a and the circumferentially-extending outside
edged portion of each of inlet port 16b and discharge port 17b can
be ensured in such a manner as to be substantially equal to the
radial distance L1. This contributes to the enhanced sealing
performance.
[0231] In contrast, the radial width (the radial wall thickness) L1
of cam-ring cylindrical portion 5a, overlapping with the
end-of-drawing-in-action region (see a part of cam-ring cylindrical
portion 5a near pump chambers r3, r3 in FIG. 5) of the inlet
portion (inlet port 16b) and the beginning-of-discharging-action
region (see a part of cam-ring cylindrical portion 5a near pump
chamber r4 in FIG. 5) of the discharge portion (discharge port
17b), is dimensioned to be shorter than the radial width (the
radial wall thickness) L2.
[0232] The provision of the thin-walled part contributes to
light-weight of cam-ring 5. That is, when the fluid pressure
(discharge pressure Pd) in the discharge portion develops
sufficiently at high engine speeds (at high revolution speeds of
pump VP), cam ring 5 tends to oscillate or rotate in the direction
that the eccentricity |OP| of cam ring 5 reduces. Hence, regarding
the thin-walled part of cam ring 5, the radial distance between
cam-ring outer peripheral surface 50a and the
circumferentially-extending outside edged portion of each of inlet
port 16b and discharge port 17b becomes smaller than the radial
distance L1. On the other hand, the rate of change of the volume of
each of pump chambers r1-r7, obtained in the high engine speed
range, becomes low. Hence, the pressure difference between inlet
pressure in inlet port 16b and discharge pressure Pd in discharge
port 17b becomes less, and thus there is a less risk of
working-fluid leakage. That is, it is unnecessary to enhance the
sealing performance between the inner and outer peripheries of cam
ring 5, in particular, at the end-of-drawing-in-action region of
the inlet portion (inlet port 16b) and at the
beginning-of-discharging-action region of the discharge portion
(discharge port 17b). Accordingly, cam ring 5 can be formed or
configured to have a comparatively thin-walled light-weight part at
the end-of-drawing-in-action region of the inlet portion (inlet
port 16b) and at the beginning-of-discharging-action region of the
discharge portion (discharge port 17b).
[0233] (13) The sidewalls (pump housing 1 as well as pump cover 2)
are made of an aluminum alloy material, whereas cam ring 5 is made
of an iron-based material. Because of the sidewalls, between which
cam ring 5 is installed and which are made of an aluminum alloy
material, cam ring 5, having a required mechanical strength and a
desired shape (a desired geometry) and dimensions, can be
accurately machined. The sidewalls made of an aluminum alloy
material also realize lightening of pump VP. For instance, when cam
ring 5 is made of an iron-based sintered alloy material, cam ring
5, having a required mechanical strength and a desired shape (a
desired geometry) and dimensions, can be accurately inexpensively
machined.
[0234] (14) Cam ring 5 has a through hole (pivot bore 52) into
which a pin (pivot pin 9), serving as the fulcrum "Q" of
oscillating motion of cam ring 5, is inserted. For example, one way
to provide the fulcrum "Q" of oscillating motion of cam ring 5
without forming the through hole (pivot bore 52) is to pivotably
support a pivot pin 9 by a pair of pin-support grooves formed in
both of the inner periphery of pump housing 1 and the outer
periphery of cam ring 5, such that both axial ends of pivot pin 9
are received by the respective pin-support grooves. However, in the
case of such a pivot-pin supporting structure, during oscillating
motion of cam ring 5, there is a possibility that cam ring 5 falls
out of pivot pin 9 or there is a possibility that pivot pin 9 falls
out of the pin-support groove pair, by the force of impact. In
contrast, in the first embodiment, pivot hole 52 is formed in cam
ring 5 as a through hole, and pivot pin 9 is fitted into pivot hole
52. Hence, pivot pin 9 can be supported by pivot hole 52 around the
entire circumference of pivot pin 9. Thus, cam ring 5 can be more
certainly supported on the fulcrum "Q" of oscillating motion.
[0235] (15) Working fluid, which is discharged from the discharge
portion, is lubricating oil, which lubricating oil is supplied to
moving and/or sliding engine parts of the internal combustion
engine, the working fluid is also used as a power source for a
variable valve actuation system (a variable valve timing control
(VTC) system configured to vary a valve characteristic of the
internal combustion engine. That is, the engine employs the
hydraulically-operated variable valve actuation system (the
hydraulically-operated VTC system). Pump VP is configured to supply
discharge pressure Pd to the VTC system as well as the moving
and/or sliding engine parts. The vane type of pump VP can output a
predetermined pressure level of discharge pressure Pd from a low
pump revolution speed range (a low engine speed range). Hence, it
is possible to enhance the operation responsiveness of the variable
valve actuation system even in the low engine speed range.
Furthermore, the pump discharge capacity of pump VP is variable.
Thus, it is possible to reduce a power loss (an energy waste) by
reducing the pump discharge capacity in the high engine speed
range.
[0236] (16) Biasing member 8 is comprised of a first biasing member
(first coil spring 8a) that permanently forces cam ring 5, and a
second biasing member (second coil spring 8b) that exerts a biasing
force on cam ring 5 only when cam ring 5 oscillates a predetermined
distance, which is greater than or equal to a predetermined angular
displacement. When cam ring 5, oscillating in the direction that
the eccentricity |OP| reduces, reaches a specified oscillated
position (the holding position) at which the biasing force,
produced by the second biasing member (second coil spring 8b) is
further added to a biasing force of the first biasing member (first
coil spring 8a), a rapid rise in the biasing force of biasing
member 8 occurs. In this specified oscillated position (the holding
position), even in the presence of a rise in the oscillating force
(moment Ta-Tb), resulting from discharge pressure Pd applied on
cam-ring inner peripheral surface 50, the oscillating-force rise
can be canceled by the biasing force of biasing member 8. In more
detail, the clockwise moment Ta-Tb, resulting from discharge
pressure Pd, is balanced to the counterclockwise moment Ts,
resulting from the summed biasing force of the first and second
biasing members. Hence, oscillating motion of cam ring 5 can be
suppressed. That is, at the specified oscillated position (the
holding position), even when a rise in discharge pressure Pd occurs
due to a pump revolution speed rise, oscillating motion of cam ring
5 can be suppressed. Thus, cam ring 5 can be held at the specified
oscillated position (the holding position), until the oscillating
force, resulting from discharge pressure Pd, exceeds the summed
biasing force of the first and second biasing members (first and
second coil springs 8a-8b) due to a further rise in discharge
pressure Pd. Therefore, in such an engine speed range Ne.sub.3-4, a
change (a decrease) in the pump discharge capacity can be
suppressed. As discussed above, the development of oscillating
motion of cam ring 5 and the suppression of oscillating motion of
cam ring 5 can be finely precisely controlled depending on the pump
revolution speed (discharge pressure Pd). Accordingly, it is
possible to realize or achieve a plurality of different pump
discharge capacities (a plurality of discharge pressure
characteristics), suited to respective pump revolution speed ranges
(respective engine speed ranges Ne.sub.1-2, Ne.sub.2-3, Ne.sub.3-4,
Ne.sub.4-5).
[0237] (17) Concretely, biasing member 8 is constructed by a
plurality of springs (first and second coil springs 8a-8b), and cam
ring 5 is forced only by one of the plurality of springs when an
oscillated amount of cam ring 5 is less than or equal to a
predetermined threshold value, and cam ring 5 is forced by the
plurality of springs (first and second coil springs 8a-8b) when the
oscillated amount of cam ring 5 exceeds the predetermined threshold
value. Thus, it is possible to certainly provide the operation and
effects of the above-mentioned item (16).
[0238] Additionally, biasing member 8 is configured such that a
biasing force (a spring constant) per unit oscillated amount
increases, as an oscillated amount of cam ring 5 increases.
[0239] In the first embodiment, to provide a nonlinear
characteristic of biasing member 8, two different resilient or
elastic members (first and second coil springs 8a-8b) are used.
Three or more resilient members (two or more elastic members) may
be used. Furthermore, in the first embodiment, a coil spring is
used as biasing member 8. Another type of spring, such as a torsion
spring or a coned disc spring, may be used as biasing member 8. In
the shown embodiment, the coiled compression spring made up of an
elastic metal, such as steel, wound into a coil, is used. Instead
of using such a metal spring, a rubber spring may be used.
Additionally, biasing member 8 is not limited to a compression
spring. An extension spring may be used. Moreover, in the shown
embodiment, to provide a nonlinear characteristic (a nonlinear
spring rate), a plurality of coil springs (first and second coil
springs 8a-8b) are combined. In lieu thereof, a single coiled
spring having a variable pitch and wire diameter may be used. For
example, a tapered outside diameter and pitch coil spring may be
used. The tapered outside diameter and pitch coil spring will
provide a more compact spring design. Alternatively, a variable
pitch spring (such as a tapered pitch spring) may be used. The
variable pitch spring is superior in reduced or suppressed costs,
as compared to a double spring structure.
[0240] (18) Working fluid, which is discharged from the discharge
portion, is supplied to a variable valve timing control (VTC)
system of the internal combustion engine. The VTC system is
configured to hold an engine valve timing at a locked state during
a startup period of the engine, and further configured to release
the locked state of the valve timing by a pressure of the working
fluid discharged from the discharge portion after the engine has
been started up, so as to permit the valve timing to be varied to a
desired valve timing. A pressure level P1 of the working-fluid
pressure, at which the locked state of the valve timing is
released, is set to be lower than a pressure level P2 of the
working-fluid pressure, at which cam ring 5 begins to operate
against a biasing force of biasing member 8. That is, it is
possible to effectively reduce a power loss (an energy waste) by
bringing the pump discharge characteristic closer to the minimum
required hydraulic characteristic curve, utilizing the nonlinear
characteristic of biasing member 8. When releasing the locked state
after the engine has been started up, it is possible to enable the
locked state to be released under a preferable state where the pump
discharge capacity becomes maximum immediately before cam ring 5
begins to operate (or oscillate) from its initial position. This
contributes to the improved operation responsiveness of the VTC
system.
[0241] (19) The VTC system is configured to operate by the pressure
of the working fluid, discharged from the discharge portion, and
further configured to be able to operate in a state where cam ring
5 is forced only by one (first coil spring 8a) of the plurality of
springs. Thus, it is possible to effectively reduce a power loss
(an energy waste) by bringing the pump discharge characteristic
closer to the minimum required hydraulic characteristic curve,
utilizing the nonlinear characteristic of biasing member 8. When
operating the VTC system by discharge pressure Pd from the
discharge portion, it is possible to always produce hydraulic
pressure required to operate the VTC system, even in the specific
state where only one spring (first coil spring 8a) forces cam ring
5. This contributes to the improved operation responsiveness of the
VTC system.
Second Embodiment
Construction of Vane Pump of Second Embodiment
[0242] Referring now to FIGS. 18-23, there is shown the variable
displacement vane pump VP of the second embodiment. The
construction of pump VP of the second embodiment is similar to that
of the first embodiment, except that the installation position of
biasing member 8, used to force cam ring 5 toward its initial
position, is changed to the third quadrant of the orthogonal
coordinate system. That is, except for the installation position of
biasing member 8, the vane-pump construction of the second
embodiment shown in FIGS. 18-23, is similar to the vane-pump
construction of the first embodiment shown in respective FIGS. 3-6
and 10-11. Thus, the same reference signs used to designate
elements in pump VP of the first embodiment shown in FIGS. 3-6 and
10-11 will be applied to the corresponding elements used in the
second embodiment shown in FIGS. 18-23, for the purpose of
comparison of the first and second embodiments. The different
layout of biasing member 8 will be hereinafter described in detail
with reference to the accompanying drawings, while detailed
description of the other pump construction will be omitted because
the above description thereon seems to be self-explanatory.
[0243] As seen in FIG. 18, in the vane pump system of the second
embodiment, biasing member 8, having the same double spring
structure (first and second coil springs 8a-8b) as the first
embodiment, is installed coaxially in a spring chamber 19 formed in
pump housing 1. Biasing member 8 forces cylindrical portion 5a of
cam ring 5 in one direction by a biasing force, so as to produce a
moment by which cam ring 5 can be rotated about pivot pin 9.
Biasing member 8 permanently forces cam ring 5 in the
maximum-eccentricity direction that the geometric center "P" of
cam-ring inner peripheral surface 50 and the rotation center "O" of
rotor 4 are spaced apart from each other.
[0244] As seen in FIG. 19, a third swelling portion 1d is formed
integral with peripheral wall 13 of pump housing 1 and located
between female screw-threaded portions 14e-14f in such a manner as
to swell radially outwards from pump-housing cylindrical portion 1a
in a combined direction of the negative x-axis direction and the
negative y-axis direction. That is, third swelling portion 1d is
laid out within the third quadrant of the orthogonal coordinate
system, which third quadrant is defined as {(x, y)|x<0, y<0}.
Third swelling portion 1d is formed as a hollow rectangular
parallelopiped. Third swelling portion 1d has spring chamber 19
formed therein.
[0245] As viewed from the z-axis direction, the inner peripheral
surface of spring chamber 19 is formed into a substantially
rectangular recessed shape. Spring chamber 19 is configured to be
surrounded in three directions by two parallel wall surfaces
19b-19c, extending in the radial direction of cylindrical portion
1a, and a bottom face 19a formed to be substantially perpendicular
to these surfaces 19b-19c. Spring chamber 19 is configured to open
on inner peripheral surface 13a of peripheral wall 13 of
cylindrical portion 1a. Two shoulder portions (engaging portions)
19d-19e, extending in the circumferential direction of cylindrical
portion 1a and opposed to each other in the circumferential
direction, are formed on peripheral wall 13 at the opening end of
spring chamber 19.
[0246] As discussed above, in the second embodiment, third swelling
portion 1d is formed with spring chamber 19, whereas second
swelling portion 1c is not formed with any spring chamber. The
dimension of the non-spring-chamber equipped second swelling
portion 1c shown in FIG. 19, measured in the x-axis direction, is
somewhat less than that of the spring-chamber equipped second
swelling portion 1c shown in FIG. 4. As viewed from the z-axis
direction in FIG. 19, the inner peripheral surface of second
swelling portion 1c is formed into a substantially rectangular
recessed shape. Second swelling portion 1c is configured to be
surrounded in three directions by two parallel wall surfaces
15j-15k, extending from cylindrical portion 1a in the positive
x-axis direction, and a wall surface 151 arranged parallel to the
y-axis. Second swelling portion 1c is configured to open on the
inner peripheral surface 13a of peripheral wall 13 of cylindrical
portion 1a. Inlet hole 16a is arranged to bestride the boundary
between the rightmost end of cylindrical portion 1a and the
leftmost end of second swelling portion 1c.
[0247] The distance from wall surface 151, arranged on the side of
second swelling portion 1c of the second embodiment (see FIG. 19)
in the positive x-axis direction, to inlet hole 16a is dimensioned
to be shorter than the distance from wall surface 15c (or wall
surface 15g), arranged on the side of second swelling portion 1c of
the first embodiment (see FIG. 4) in the positive x-axis direction,
to inlet hole 16a. As viewed from the z-axis direction, in the
first embodiment (see FIG. 4), the peripheral portion 13b and
shoulder portion (engaging portion) 15h are configured to overlap
with inlet hole 16a. In contrast, in the second embodiment (see
FIG. 19), second swelling portion 1c is not formed with any spring
chamber, and thus pump housing 1 does not have any obstruction that
impedes the flow of working fluid from inlet hole 16a into pump
housing 1.
[0248] As viewed in the direction of the axis "O", a cam-ring
receiving portion 13c is formed on the inner peripheral surface 13a
of cylindrical portion 1a and arranged on the side of cylindrical
portion 1a in the positive y-axis, in such a manner as to slightly
protrude radially inwards. A moderately curved concave stopper
surface 13d is formed on cam-ring receiving portion 13c, while
facing in the negative y-axis direction. As viewed from the z-axis
direction, stopper surface 13d is formed into a circular-arc shape,
which is substantially conformable to the shape (the curvature) of
the outer peripheral surface of cam ring 5.
[0249] FIG. 20 shows the initial setting state of cam ring 5.
Cam-ring cylindrical portion 5a has a radially-protruding portion
5e formed integral with its outer periphery. The protruding portion
5e is formed to have the same length as cylindrical portion 5a in
the z-axis direction. Protruding portion 5e is laid out on the side
of the fulcrum "Q" of oscillating motion (pivot portion 5c) with
respect to the geometric center "P" of cam-ring cylindrical portion
5a, and laid out on the side of the negative y-axis direction with
respect to the fulcrum "Q" of oscillating motion. That is,
protruding portion 5e is laid out within the third quadrant of the
orthogonal coordinate system, which third quadrant is defined as
{(x, y)|x<0, y<0}, in the same manner as third swelling
portion 1d having spring chamber 19. The surface 50e of protruding
portion 5e is formed as a curved surface. As viewed from the z-axis
direction, protruding portion 5e is formed into a semicircle in
cross section. The root (the basal end) of protruding portion 5e is
contoured to be continuous with the outer peripheral surface of
cylindrical portion 5a.
[0250] In the initial setting state of cam ring 5, as viewed in the
radial direction of cylindrical portion 1a, the tip of protruding
portion 5e is laid out substantially at the same position as the
midpoint of two shoulder portions (engaging portions) 19d-19e of
spring chamber 19 opposed to each other in the circumferential
direction of cylindrical portion 1a. As viewed in the
circumferential direction of cylindrical portion 1a, the centerline
of protruding portion 5e is aligned with the centerline of spring
chamber 19. The maximum width of protruding portion 5e is
dimensioned to be less than the maximum width of the opening of
spring chamber 19, in other words, the distance between the opposed
shoulder portions 19d-19e. In the initial setting state of FIG. 20,
the outer peripheral surface of the positive y-axis direction of
cylindrical portion 5a (the uppermost end of cylindrical portion
5a, viewing FIG. 20)) is brought into wall-contact with stopper
surface 13d of cam-ring receiving portion 13c. That is, in the
initial setting state, cam-ring cylindrical portion 5a seats on
stopper surface 13d.
[0251] FIG. 21 shows the cross section of pump VP whose pump cover
is installed, taken along the line F-F of FIG. 20. The dimensional
relationship between biasing member 8 and spring chamber 19 in the
second embodiment is identical to that between biasing member 8 and
spring chamber 15d in the first embodiment (see FIG. 6). In a
similar manner to the first embodiment, first and second coil
springs 8a-8b, constructing biasing member 8, are installed in
spring chamber 19. First coil spring 8a is installed in spring
chamber 19 and disposed between a spring-chamber bottom face 19a
and cam-ring protruding portion 5e under a preloaded condition
where first coil spring 8a is preloaded by an initial set load W1'.
Second coil spring 8b is also installed in spring chamber 19 and
disposed between spring-chamber bottom face 19a and the shoulder
pair 19d-19e under a preloaded condition where second coil spring
8b is preloaded by an initial set load W3'.
[0252] In a similar manner to FIG. 10 showing the holding state of
cam ring 5 of pump VP of the first embodiment, FIG. 22 shows the
cam-ring holding state of pump VP of the second embodiment, where
the tip of protruding portion 5e is brought into abutted-engagement
with the diametrically-opposing coil-end portions (opposed to each
other in the z-axis direction) of the upper coil end of second coil
spring 8b, facing radially inwards, and thus the eccentricity |OP|
of geometric center "P" of cam-ring inner peripheral surface 50
with respect to the axis "O" of drive shaft 3 becomes a
substantially intermediate value between the maximum and minimum
eccentricities. In a similar manner to FIG. 11 showing the
minimum-eccentricity state of cam ring 5 of pump VP of the first
embodiment, FIG. 23 shows the cam-ring minimum-eccentricity state
of pump VP of the second embodiment, where the eccentricity |OP| of
geometric center "P" of cam-ring inner peripheral surface 50 with
respect to the axis "O" of drive shaft 3 becomes minimum (zero),
that is, the axis "O" and the geometric center "P" accord with each
other. In the minimum-eccentricity state, the left-hand side
surface (viewing FIG. 23) of protruding portion 5, facing pivot
portion 5c, is brought into abutted-engagement with the left-hand
side shoulder portion 19e. Under these conditions, the shoulder
portion 19e serves as a stopper for restricting or preventing a
further clockwise rotation of cam ring 5 about the fulcrum "Q".
Operation of Second Embodiment
[0253] Force Fs (a biasing force), produced by biasing member 8,
acts on cam ring 5. In the second embodiment, force Fs corresponds
to a force vector acting on protruding portion 5e of cam-ring
cylindrical portion 5a in a combined direction of the positive
x-axis direction and the positive y-axis direction. Force Fs
produces a moment Ts by which cam ring 5 can be rotated about the
fulcrum "Q" of oscillating motion in the counterclockwise direction
that the eccentricity |OP| of geometric center "P" of cam-ring
inner peripheral surface 50 with respect to the axis "O" of drive
shaft 3 (or rotor 4) increases. The distance from the fulcrum "Q"
to the point of application of force Fs (i.e., protruding portion
5e) in pump VP of the second embodiment is dimensioned to be
shorter than the distance from the fulcrum "Q" to the point of
application of force Fs (i.e., protruding portion 54 of cam-ring
arm portion 5d) in pump VP of the first embodiment. Hence, to
produce the same magnitude of moment Ts as the first embodiment, in
the case of the second embodiment, the biasing force (the spring
load) of biasing member 8 has to be set to a larger value. However,
as described previously, by appropriately setting the position of
the fulcrum "Q" of oscillating motion of cam ring 5, the cam-ring
oscillating force, resulting from discharge pressure Pd, can be set
to a smaller value. As a result, it is possible to set the spring
force, produced by biasing member 8 against the cam-ring
oscillating force resulting from discharge pressure Pd, to a
smaller value. Accordingly, it is unnecessary to use a bigger
biasing member 8 in the vane pump system of the second embodiment,
as compared to the size of biasing member 8 used in pump VP of the
first embodiment.
[0254] As previously discussed, in the second embodiment, the
installation position of biasing member 8 and the location of
spring chamber 19 are changed from second swelling portion 1c to
third swelling portion 1d. Hence, the size of second swelling
portion 1c can be suppressed or downsized to a minimum size that
satisfies a space requirement for inlet hole 16a. Additionally,
pump housing 1 does not have any obstruction that impedes the flow
of working fluid from inlet hole 16a into pump housing 1. This
enables a more smooth flow of working fluid from inlet hole 16a
toward inlet port 16b or toward the pump chambers located on the
pump inlet side, thus enhancing the pump suction efficiency. In the
second embodiment, by optimizing the layout of biasing member 8 (or
spring chamber 19) as set forth above, it is possible to reconcile
or balance two contradictory requirements, that is, the
compactly-designed vane pump VP and the enhanced pump efficiency,
while providing the operation and effects as the first
embodiment.
Effects of Second Embodiment
[0255] (20) Biasing member 8 is located outside of an outer
periphery of cam ring 5 and laid out to be offset toward the
fulcrum "Q" of oscillating motion with respect to the geometric
center "P" of cam-ring inner peripheral surface 50. In this manner,
by optimizing the layout of biasing member 8 (or spring chamber 19)
with respect to the location of inlet hole 16a (or inlet port 16b),
it is possible to enhance the pump suction efficiency, while
realizing the compactly-designed vane pump VP.
Third Embodiment
[0256] Referring now to FIGS. 24-25, there is shown the variable
displacement vane pump VP of the third embodiment. The construction
of pump VP of the third embodiment is similar to that of the first
embodiment, except that the shape of cam ring 5 and the location
and shape of discharge port 17b of the third embodiment differ from
those of the first embodiment. Concretely, in the first embodiment
cam ring 5 is formed with communication hole 51, whereas in the
third embodiment cam ring 5 is not formed with such a communication
hole. The other pump construction of the third embodiment shown in
FIGS. 24-25 is similar to the vane-pump construction of the first
embodiment shown in respective FIGS. 4-5. Thus, the same reference
signs used to designate elements in pump VP of the first embodiment
shown in FIGS. 4-5 will be applied to the corresponding elements
used in the third embodiment shown in FIGS. 24-25, for the purpose
of comparison of the first and third embodiments. The different
shape of cam ring 5 and the location and shape of discharge port
17b will be hereinafter described in detail with reference to the
accompanying drawings, while detailed description of the other pump
construction will be omitted because the above description thereon
seems to be self-explanatory.
[0257] In a similar manner to FIG. 4 showing the front elevation
view of pump VP of the first embodiment, FIG. 24 shows the
structure of pump housing 1 of pump VP of the third embodiment. In
the third embodiment, discharge port 17b has only the circular-arc
shaped groove 17c arranged on bottom face 10a of cylindrical
portion 1a, but not have sector groove 17d. First swelling portion
1b is formed to slightly swell radially outwards from pump-housing
cylindrical portion 1a so as to be able to accommodate therein only
the cam-ring pivot portion 5c. Pump cover 2 is configured to be
substantially conformable to the shape of pump housing 1 having the
slightly radially-outward extending first swelling portion 1b.
[0258] Discharge hole 17a is not located within first swelling
portion 1b, but formed in bottom face 10a of pump-housing basal
portion 10 in such a manner as to be laid out on the line segment
linking the fulcrum "Q" of oscillating motion and the axis "O" of
drive shaft 3 (or the rotation center of rotor 4). As viewed from
the z-axis direction, discharge hole 17a is laid out on the side of
the positive x-axis direction with respect to support portion 12a,
in such a manner as to overlap with all of discharge port 17b
(circular-arc shaped groove 17c), oil storage portion 18a, and
bearing lubrication oil groove 18d. As viewed from the z-axis
direction, discharge hole 17a is configured to open into only the
discharge port 17b (circular-arc shaped groove 17c). Discharge hole
17a is communicated with the inside of pump housing 1 through
discharge port 17b (circular-arc shaped groove 17c).
[0259] In a similar manner to FIG. 5 showing the initial setting
state of cam ring 5 of pump VP of the first embodiment, FIG. 25
shows the cam-ring initial setting state of pump VP of the third
embodiment. Cam ring 5 of pump VP of the third embodiment does not
have sector portion 5b and communication hole 51. Cam ring 5 is
configured to oscillate or rotate about pivot portion 5c (the
fulcrum "Q" of oscillating motion) installed in first swelling
portion 1b. By rotary motion of rotor 4, discharge pressure Pd is
supplied from the pump chambers, defined on the side of the
discharge portion, through discharge port 17b (circular-arc shaped
groove 17c) into discharge hole 17a.
Effects of Third Embodiment
[0260] (21) As viewed from the z-axis direction, discharge hole 17a
is laid out at a specified position where discharge hole 17a
overlaps with discharge port 17b (circular-arc shaped groove 17c).
In other words, as viewed from the z-axis direction, discharge hole
17a is laid out at a specified position where discharge hole 17a
overlaps with the cam-ring inner peripheral surface 50 (the pump
chambers), with which the tips of vanes 6a-6g are in
sliding-contact.
[0261] That is to say, in the case of the vane pump construction of
the first embodiment, in which discharge hole 17a is arranged
radially outside of circular-arc shaped groove 17c, a working-fluid
passage (i.e., sector groove 17d), intercommunicating circular-arc
shaped groove 17c and discharge hole 17a, has to be formed in pump
housing 1. For the purpose of increasing the number of fluid
passages, required to communicate each of pump-housing side
circular-arc shaped groove 17c and pump-cover side circular-arc
shaped groove 23 with discharge hole 17a, in the first embodiment,
cam ring 5 is formed with communication hole 51, and pump cover 2
is formed with sector groove 23d. In contrast, the layout of
discharge hole 17a of pump VP of the third embodiment eliminates
the necessity of communication hole 51 and pump-cover sector groove
23d. Additionally, cam ring 5 can be dimensioned or set to a
minimum required size, thus realizing a compact vane pump
design.
Fourth Embodiment
[0262] Referring now to FIG. 26, there is shown the variable
displacement vane pump VP of the fourth embodiment. Basically, the
construction of pump VP of the fourth embodiment is similar to that
of the first embodiment. However, the vane pump construction of the
fourth embodiment somewhat differs from that of the first
embodiment, in that the pump construction of the fourth embodiment
is realized by combining the layout of biasing member 8 (and spring
chamber 19) of the pump of the second embodiment (see FIGS. 18-23)
with the layout of discharge hole 17a (the shape of and location of
discharge port 17b and the shape of cam ring 5) of the pump of the
third embodiment (see FIGS. 24-25). Hence, the fourth embodiment
can provide the same effects as the above-mentioned items
(20)-(21).
[0263] In the shown embodiments, variable displacement vane pump VP
is applied to an internal combustion engine of an automotive
vehicle. Variable displacement vane pump VP may be applied to
another type of machineries, such as a hydraulically-operated
crane.
[0264] In the shown embodiments, pump VP is used for supplying
moving engine parts with lubricating oil and for delivering oil
(serving as a working medium as well as a lubricating substance) to
a variable valve actuation mechanism. Pump VP may be used as a
drive source (a power steering pump) for a hydraulic power steering
system.
[0265] In the shown embodiments, pump VP is driven by an internal
combustion engine. Pump VP may be driven by another type of driving
power source, such as an electric motor. Also, the vane rotor of
pump VP is driven in synchronism with rotation of a crankshaft of
an internal combustion engine. It is not necessary that rotation of
the vane rotor should be synchronized with rotation of the
crankshaft.
[0266] In the first embodiment, as a variable valve actuation
system that utilizes discharge pressure Pd from pump VP, a variable
valve timing control (VTC) system is used. Discharge pressure Pd
from pump VP may be utilized for another type of variable valve
actuation system, for example, a hydraulically-operated variable
valve lift (VVL) system configured to variably control a valve lift
and valve timing.
[0267] In the first embodiment, the VTC system is applied to only
the intake-valve side of the engine. In lieu thereof, the VTC
system may be applied to at least one of the intake-valve side and
the exhaust-valve side of the engine.
[0268] In the first embodiment, a first group of grooves (inlet
port 16b and discharge port 17b) are formed in pump housing 1,
whereas a second group of grooves (inlet port 22 and discharge port
23) are formed in pump cover 2. For the purpose of more simplified
pump construction, reduced machining time and costs, and lower pump
system costs, grooves (a pump inlet port and a pump discharge port)
may be formed only in either pump housing 1 or pump cover 2.
[0269] In the first embodiment, any seal member is not interleaved
between pump-housing flanged portion 14 and pump-cover flanged
portion 24. In order to certainly prevent oil leakage from the
inside of pump VP and to enhance a fluid-tight performance of pump
VP, a seal member may be interleaved between them.
[0270] In the first embodiment, vanes 6a-6g, each of which is
fitted into rotor 4 to slide from rotor 4 toward inner peripheral
surface 50 of cam ring 5 and set to be kept in abutted-engagement
with cam-ring inner peripheral surface 50 by means of vane rings
7a-7b. It is not necessary that vanes 6a-6g should be kept in
abutted-engagement with cam-ring inner peripheral surface 50. A
slight clearance space defined between the tip of each of vanes
6a-6g and cam-ring inner peripheral surface 50 may be permitted to
the extent that noise, caused by collision-contact between the tip
of each of vanes 6a-6g with cam-ring inner peripheral surface 50
owing to radially-outward movements of vanes 6a-6g out of
respective slits 4a-4g, is negligible at the beginning of rotation
of pump VP.
[0271] In the first and third embodiments, biasing member 8 (or
spring chamber 15d) is arranged outside of cam ring 5 and laid out
at the position substantially symmetrical to the fulcrum "Q" of
oscillating motion of cam ring 5 with respect to the rotation
center "O" of vane rotor 4 (or with respect to the geometric center
"P" of cam-ring inner peripheral surface 50). In contrast, in the
second and fourth embodiments, biasing member 8 (or spring chamber
19) is arranged outside of cam ring 5 and laid out on the side of
the fulcrum "Q" of oscillating motion of cam ring 5 with respect to
the rotation center "O" of vane rotor 4 (or with respect to the
geometric center "P" of cam-ring inner peripheral surface 50). That
is, biasing member 8 may be laid out at an arbitrary
circumferential position at which cam ring 5 can be forced in the
direction that the volume difference between the volume of the
largest pump chamber (pump chamber r4 in FIG. 5) of pump chambers
r1-r7 and the volume of the smallest pump chamber (pump chambers
r1, r7 in FIG. 5) increases.
[0272] In the shown embodiments, cam ring 5 is formed with an pin
insertion hole (pivot bore 52) into which a pin (pivot pin 9),
serving as a fulcrum of oscillating motion of cam ring 5, is
fitted. In lieu thereof, pivot pin 9 may be pivotably supported by
a pair of pin-support grooves formed in both of the inner periphery
of pump housing 1 and the outer periphery of cam ring 5, such that
both axial ends of pivot pin 9 are received by the respective
pin-support grooves.
[0273] The entire contents of Japanese Patent Application No.
2008-133889 (filed May 22, 2008) are incorporated herein by
reference.
[0274] While the foregoing is a description of the preferred
embodiments carried out the invention, it will be understood that
the invention is not limited to the particular embodiments shown
and described herein, but that various changes and modifications
may be made without departing from the scope or spirit of this
invention as defined by the following claims.
* * * * *